US4967557A - Control system for load-sensing hydraulic drive circuit - Google Patents

Control system for load-sensing hydraulic drive circuit Download PDF

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Publication number
US4967557A
US4967557A US07/301,718 US30171889A US4967557A US 4967557 A US4967557 A US 4967557A US 30171889 A US30171889 A US 30171889A US 4967557 A US4967557 A US 4967557A
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Prior art keywords
delivery amount
pressure
pump
target delivery
differential pressure
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US07/301,718
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Inventor
Eiki Izumi
Yasuo Tanaka
Hiroshi Watanabe
Kuniaki Yoshida
Toichi Hirata
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Hitachi Construction Machinery Co Ltd
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Hitachi Construction Machinery Co Ltd
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Assigned to HITACHI CONSTRUCTION MACHINERY CO., LTD., A CORP. OF JAPAN reassignment HITACHI CONSTRUCTION MACHINERY CO., LTD., A CORP. OF JAPAN ASSIGNMENT OF ASSIGNORS INTEREST. Assignors: HIRATA, TOICHI, IZUMI, EIKI, TANAKA, YASUO, WATANABE, HIROSHI, YOSHIDA, KUNIAKI
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B9/00—Servomotors with follow-up action, e.g. obtained by feed-back control, i.e. in which the position of the actuated member conforms with that of the controlling member
    • F15B9/02—Servomotors with follow-up action, e.g. obtained by feed-back control, i.e. in which the position of the actuated member conforms with that of the controlling member with servomotors of the reciprocatable or oscillatable type
    • F15B9/08—Servomotors with follow-up action, e.g. obtained by feed-back control, i.e. in which the position of the actuated member conforms with that of the controlling member with servomotors of the reciprocatable or oscillatable type controlled by valves affecting the fluid feed or the fluid outlet of the servomotor
    • F15B9/10—Servomotors with follow-up action, e.g. obtained by feed-back control, i.e. in which the position of the actuated member conforms with that of the controlling member with servomotors of the reciprocatable or oscillatable type controlled by valves affecting the fluid feed or the fluid outlet of the servomotor in which the controlling element and the servomotor each controls a separate member, these members influencing different fluid passages or the same passage
    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02F—DREDGING; SOIL-SHIFTING
    • E02F9/00—Component parts of dredgers or soil-shifting machines, not restricted to one of the kinds covered by groups E02F3/00 - E02F7/00
    • E02F9/20—Drives; Control devices
    • E02F9/2025—Particular purposes of control systems not otherwise provided for
    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02F—DREDGING; SOIL-SHIFTING
    • E02F9/00—Component parts of dredgers or soil-shifting machines, not restricted to one of the kinds covered by groups E02F3/00 - E02F7/00
    • E02F9/20—Drives; Control devices
    • E02F9/22—Hydraulic or pneumatic drives
    • E02F9/2221—Control of flow rate; Load sensing arrangements
    • E02F9/2225—Control of flow rate; Load sensing arrangements using pressure-compensating valves
    • E02F9/2228—Control of flow rate; Load sensing arrangements using pressure-compensating valves including an electronic controller
    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02F—DREDGING; SOIL-SHIFTING
    • E02F9/00—Component parts of dredgers or soil-shifting machines, not restricted to one of the kinds covered by groups E02F3/00 - E02F7/00
    • E02F9/20—Drives; Control devices
    • E02F9/22—Hydraulic or pneumatic drives
    • E02F9/2221—Control of flow rate; Load sensing arrangements
    • E02F9/2232—Control of flow rate; Load sensing arrangements using one or more variable displacement pumps
    • E02F9/2235—Control of flow rate; Load sensing arrangements using one or more variable displacement pumps including an electronic controller
    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02F—DREDGING; SOIL-SHIFTING
    • E02F9/00—Component parts of dredgers or soil-shifting machines, not restricted to one of the kinds covered by groups E02F3/00 - E02F7/00
    • E02F9/20—Drives; Control devices
    • E02F9/22—Hydraulic or pneumatic drives
    • E02F9/2246—Control of prime movers, e.g. depending on the hydraulic load of work tools
    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02F—DREDGING; SOIL-SHIFTING
    • E02F9/00—Component parts of dredgers or soil-shifting machines, not restricted to one of the kinds covered by groups E02F3/00 - E02F7/00
    • E02F9/20—Drives; Control devices
    • E02F9/22—Hydraulic or pneumatic drives
    • E02F9/2278—Hydraulic circuits
    • E02F9/2292—Systems with two or more pumps
    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02F—DREDGING; SOIL-SHIFTING
    • E02F9/00—Component parts of dredgers or soil-shifting machines, not restricted to one of the kinds covered by groups E02F3/00 - E02F7/00
    • E02F9/20—Drives; Control devices
    • E02F9/22—Hydraulic or pneumatic drives
    • E02F9/2278—Hydraulic circuits
    • E02F9/2296—Systems with a variable displacement pump
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02D—CONTROLLING COMBUSTION ENGINES
    • F02D29/00—Controlling engines, such controlling being peculiar to the devices driven thereby, the devices being other than parts or accessories essential to engine operation, e.g. controlling of engines by signals external thereto
    • F02D29/04—Controlling engines, such controlling being peculiar to the devices driven thereby, the devices being other than parts or accessories essential to engine operation, e.g. controlling of engines by signals external thereto peculiar to engines driving pumps
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B11/00—Servomotor systems without provision for follow-up action; Circuits therefor
    • F15B11/16—Servomotor systems without provision for follow-up action; Circuits therefor with two or more servomotors
    • F15B11/161—Servomotor systems without provision for follow-up action; Circuits therefor with two or more servomotors with sensing of servomotor demand or load
    • F15B11/165—Servomotor systems without provision for follow-up action; Circuits therefor with two or more servomotors with sensing of servomotor demand or load for adjusting the pump output or bypass in response to demand
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B21/00—Common features of fluid actuator systems; Fluid-pressure actuator systems or details thereof, not covered by any other group of this subclass
    • F15B21/08—Servomotor systems incorporating electrically operated control means
    • F15B21/087—Control strategy, e.g. with block diagram
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/20—Fluid pressure source, e.g. accumulator or variable axial piston pump
    • F15B2211/205—Systems with pumps
    • F15B2211/2053—Type of pump
    • F15B2211/20546—Type of pump variable capacity
    • F15B2211/20553—Type of pump variable capacity with pilot circuit, e.g. for controlling a swash plate
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/20—Fluid pressure source, e.g. accumulator or variable axial piston pump
    • F15B2211/255—Flow control functions
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/30—Directional control
    • F15B2211/305—Directional control characterised by the type of valves
    • F15B2211/30505—Non-return valves, i.e. check valves
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/30—Directional control
    • F15B2211/305—Directional control characterised by the type of valves
    • F15B2211/30525—Directional control valves, e.g. 4/3-directional control valve
    • F15B2211/3053—In combination with a pressure compensating valve
    • F15B2211/30535—In combination with a pressure compensating valve the pressure compensating valve is arranged between pressure source and directional control valve
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/30—Directional control
    • F15B2211/31—Directional control characterised by the positions of the valve element
    • F15B2211/3105—Neutral or centre positions
    • F15B2211/3111—Neutral or centre positions the pump port being closed in the centre position, e.g. so-called closed centre
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/30—Directional control
    • F15B2211/315—Directional control characterised by the connections of the valve or valves in the circuit
    • F15B2211/3157—Directional control characterised by the connections of the valve or valves in the circuit being connected to a pressure source, an output member and a return line
    • F15B2211/31576—Directional control characterised by the connections of the valve or valves in the circuit being connected to a pressure source, an output member and a return line having a single pressure source and a single output member
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/30—Directional control
    • F15B2211/32—Directional control characterised by the type of actuation
    • F15B2211/329—Directional control characterised by the type of actuation actuated by fluid pressure
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/60—Circuit components or control therefor
    • F15B2211/605—Load sensing circuits
    • F15B2211/6051—Load sensing circuits having valve means between output member and the load sensing circuit
    • F15B2211/6054—Load sensing circuits having valve means between output member and the load sensing circuit using shuttle valves
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/60—Circuit components or control therefor
    • F15B2211/63—Electronic controllers
    • F15B2211/6303—Electronic controllers using input signals
    • F15B2211/6306—Electronic controllers using input signals representing a pressure
    • F15B2211/6309—Electronic controllers using input signals representing a pressure the pressure being a pressure source supply pressure
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/60—Circuit components or control therefor
    • F15B2211/63—Electronic controllers
    • F15B2211/6303—Electronic controllers using input signals
    • F15B2211/6306—Electronic controllers using input signals representing a pressure
    • F15B2211/6313—Electronic controllers using input signals representing a pressure the pressure being a load pressure
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/60—Circuit components or control therefor
    • F15B2211/63—Electronic controllers
    • F15B2211/6303—Electronic controllers using input signals
    • F15B2211/633—Electronic controllers using input signals representing a state of the prime mover, e.g. torque or rotational speed
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/60—Circuit components or control therefor
    • F15B2211/63—Electronic controllers
    • F15B2211/6303—Electronic controllers using input signals
    • F15B2211/6333—Electronic controllers using input signals representing a state of the pressure source, e.g. swash plate angle
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/60—Circuit components or control therefor
    • F15B2211/63—Electronic controllers
    • F15B2211/6303—Electronic controllers using input signals
    • F15B2211/634—Electronic controllers using input signals representing a state of a valve
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/60—Circuit components or control therefor
    • F15B2211/63—Electronic controllers
    • F15B2211/6303—Electronic controllers using input signals
    • F15B2211/6346—Electronic controllers using input signals representing a state of input means, e.g. joystick position
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/60—Circuit components or control therefor
    • F15B2211/635—Circuits providing pilot pressure to pilot pressure-controlled fluid circuit elements
    • F15B2211/6355—Circuits providing pilot pressure to pilot pressure-controlled fluid circuit elements having valve means
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/70—Output members, e.g. hydraulic motors or cylinders or control therefor
    • F15B2211/705—Output members, e.g. hydraulic motors or cylinders or control therefor characterised by the type of output members or actuators
    • F15B2211/7051—Linear output members
    • F15B2211/7053—Double-acting output members
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/70—Output members, e.g. hydraulic motors or cylinders or control therefor
    • F15B2211/71—Multiple output members, e.g. multiple hydraulic motors or cylinders

Definitions

  • the present invention relates to a load-sensing hydraulic drive circuit for hydraulic machines, such as hydraulic excavators and cranes, each equipped with a plurality of hydraulic actuators, and more particularly to a control system for a load-sensing hydraulic drive circuit, which is designed to control the flow rates of hydraulic fluid supplied to the hydraulic actuators using pressure compensated flow control valves, while holding the delivery pressure of a hydraulic pump higher by a predetermined value than the maximum load pressure among the hydraulic actuators.
  • the hydraulic drive circuit comprises a pressure compensated flow control valve connected between a hydraulic pump and each of the hydraulic actuators for controlling the flow rate of hydraulic fluid supplied to the hydraulic actuator in response to an operation signal from a control lever, and a load-sensing regulator for holding the delivery pressure of the hydraulic pump higher by a predetermined value than the maximum load pressure among the plural hydraulic actuators.
  • the pressure compensated flow control valve has a pressure compensating function to maintain the flow rate constant regardless of fluctuations in the load pressure or the delivery pressure of the hydraulic pump, so that a flow rate proportional to the operated amount of each control lever is supplied to the associated hydraulic actuator.
  • the load-sensing regulator functions to constantly maintain the delivery pressure of the hydraulic pump at a lower limit corresponding to the maximum load pressure among the hydraulic actuators for energy saving.
  • the delivery amount of a variable displacement hydraulic pump is determined by the product of its displacement, i.e., inclination angle of a swash plate, in the case of a swash plate type and the rotational speed of the pump.
  • the inclination angle of the swash plate has an upper limit determined by the pump structure, at which upper limit of the delivery amount of the pump also reaches its maximum.
  • an input torque regulator has usually been equipped on the pump to limit the maximum inclination angle of the swash plate so that input torque of the pump will not exceed output torque of the prime mover, thereby controlling the delivery amount of the pump input torque limiting control.
  • the pump cannot increase the delivery amount (inclination angle) much more even though it is under the load-sensing control. In other words, the delivery amount of the pump is saturated. As a result, the delivery pressure of the pump is reduced and can no longer be maintained higher by a predetermined value than the maximum load pressure. Thus, the delivery amount of the pump is caused to largely flow into the actuator(s) on the lower pressure side, while the hydraulic fluid is not supplied to the actuator(s) on the higher pressure side, resulting in a problem that the combined operation of plural actuators cannot be performed smoothly.
  • DE-A1-3422165 (corresponding to Japanese Patent Laid-Open No. 60-11706) has proposed such a circuit arrangement that a pair of opposing pilot chambers is added to a pressure balance valve of each pressure compensated flow control valve, and the delivery pressure of the pump is introduced to one of the pilot chambers which acts in the valve-opening direction, while the maximum load pressure among the plural actuators is introduced to the other pilot chamber which acts in the valve-closing direction.
  • the pressure compensated flow control valve determines a consumable flow rate, that is to be passed to the associated hydraulic actuator therethrough, based on both a throttle opening command value for the flow control valve given by an operation signal from the control lever and a differential pressure command value across the flow control valve given to the pressure balance valve. Both the throttle openings of the flow control valve and the pressure balance valve are controlled so that the actual flow rate through the pressure compensated flow control valve, i.e., the flow rate consumed by the actuator becomes equal to the consumable flow rate.
  • the differential pressure command value across the flow control valve is directly applied to the pressure balance valve hydraulically such that the delivery pressure of the pump and the maximum load pressure among the hydraulic actuators are introduced to the pressure balance valve in opposite directions, causing the differential pressure therebetween to act on the pressure balance valve.
  • the differential pressure command values applied to all the pressure balance valves are limited to compensate (reduce) the total consumable flow rate for all the hydraulic actuators. This reduces the total flow rate actually consumed by the actuators.
  • this type of control will be referred to as total consumable flow compensating control.
  • the differential pressure between the pump delivery pressure and the maximum load pressure is reduced responsive to deficiencies in the actual delivery pressure of the pump as compared with the demand flow rates commanded by the control levers, and hence, the total consumable flow rate is always coincident with the total of actual flow rates consumed by the hydraulic actuators.
  • the load-sensing control controls the delivery amount of the pump to hold the differential pressure constant, and has a slower response speed than that of the total consumable flow compensating control, as the control of the delivery amount of the pump is carried out through various mechanisms. Therefore, when the delivery pressure of the pump is reduced at the moment the control lever is operated to start supply of the hydraulic fluid to the actuator or increase the supply amount thereof, the flow rate through the pressure compensated flow control valve starts to be restricted under the total consumable flow compensating control before the load-sensing control starts to increase the delivery amount of the pump. This causes the problem that in a transitional period, the flow rate supplied to the actuator cannot be increased and the operability is impaired even though the control lever is operated with an intention to increase the flow rate.
  • the pump delivery amount is increased under the load-sensing control to raise up the pump delivery pressure after the flow rate through the flow control valve has been restricted under the total consumable flow compensating control, then the total consumable flow compensating control is released to increase the flow rate through the flow control valve, causing the delivery pressure of the pump to be reduced, and thereafter the flow rate through the flow control valve is restricted under the total consumable flow compensating control before the load-sensing control has started to increase the pump delivery amount.
  • the load-sensing control and the total consumable flow compensating control interfere with each other, thereby resulting in a hunting phenomenon.
  • a control system for a load-sensing hydraulic drive circuit comprising; at least one hydraulic pump; a plurality of hydraulic actuators driven with hydraulic fluid delivered from the pump; and a pressure compensated flow control valve connected between the pump and each of the actuators, for controlling a flow rate of the hydraulic fluid supplied to each actuator in response to an operation signal from control means
  • the control system comprises a first detection device for detecting a differential pressure between the delivery pressure of the pump and the maximum load pressure among the plurality of hydraulic actuators; a second detection device for detecting the delivery pressure of the pump; a first device for calculating, based on a differential pressure signal from the first detection means, a differential pressure target delivery amount Q ⁇ p of the pump to hold the differential pressure constant; a second device for calculating an input limiting target delivery amount QT of the pump based on at least a pressure signal from the second detection device an an input limiting function preset for the pump; a third device for selecting one of the differential pressure target delivery
  • the fourth device may control a pressure balance valve of the pressure compensated flow control valve based on the compensation value Qns.
  • the fourth device may calculate an operation signal modifying factor ⁇ from the compensation value Qns, modify the operation signal from the control means using the operation signal modifying factor ⁇ , and control the pressure compensated flow control valve using the corrected operation signal.
  • the third device may select smaller one of the differential pressure target delivery amount Q ⁇ p and the input limiting target delivery amount QT as the delivery amount target value Qo for the pump.
  • the third device may select the differential pressure target delivery amount Q ⁇ p as the delivery amount target value Qo for the pump when the compensation value Qns is zero, and the input limiting target delivery amount QT as the delivery amount target value Qo for the pump when the compensation value Qns is not zero.
  • the fourth device may include an adder device to determine a target delivery amount deviation ⁇ Q as a deviation between the differential pressure target delivery amount Q ⁇ p and the input limiting target delivery amount QT, and calculate the compensation value Qns using at least the target delivery amount deviation ⁇ Q.
  • the first device may include an adder device to calculate a differential pressure deviation ⁇ P' between the differential pressure signal from the first detection device and the preset target differential pressure
  • the fourth device may further include a filter device for outputting zero when the differential pressure deviation ⁇ P' is positive and a value ⁇ P" equal to the differential pressure deviation ⁇ P' when it is negative, a selector device for selecting an output ⁇ P" of the filter device when the target delivery amount deviation ⁇ Q is negative and the output ⁇ P' of the adder device when the target delivery amount deviation ⁇ Q is positive, and a calculation device for calculating the compensation value Qns from the value ⁇ P" or ⁇ P' selected by the selector device.
  • the fourth means may calculate a deviation between the compensation value Qns and a preset offset value, and then output a resulting value Qnso as the final compensation value.
  • the first device may comprise an integral type calculation device which calculates, based on the differential pressure signal from the first detection device, an increment ⁇ Q ⁇ p of the differential pressure target delivery amount Q ⁇ p for holding the differential pressure constant, and then adds the increment ⁇ Q ⁇ p to the previously calculated differential target delivery amount Qo-1 for determining the differential pressure target delivery amount Q ⁇ p;
  • second device may comprise an integral type calculation device which calculates an increment ⁇ Qps of the input limiting target delivery amount QT for controlling the pressure signal from the second detection device to a target delivery pressure Pr obtained from the input limiting function of the pump. It then adds the increment ⁇ Qps to the previously calculated input limiting target delivery amount Qo-1 for determining the input limiting target delivery amount QT.
  • the third device may comprise means for selecting one of the increment ⁇ Q ⁇ p of the differential pressure target delivery amount Q ⁇ p and the increment ⁇ Qps of the input limiting target delivery amount QT for selecting one of the differential pressure target delivery amount Q ⁇ p and the input limiting target delivery amount QT.
  • the input limiting function of the second device may be an input torque limiting function with one of the delivery pressure and the input limiting target delivery amount of the pump as a parameter, and the second device may calculate the input limiting target delivery amount QT of the pump based on both the pressure signal of the second detection device and the input torque limiting function.
  • control system may further include third detection device for determining a deviation between the target speed and the actual speed of a prime mover for driving the pump; and the input limiting function of the second device may be an input torque limiting function with one of the delivery pressure and the input limiting target delivery amount of the pump and the speed deviation of the prime mover as parameters, and the second device may calculate the input limiting target delivery amount QT of the pump based on the pressure signal of the second detection device, the speed deviation signal of the third detection device and the input torque limiting function.
  • the delivery amount of the pump is controlled such that the differential pressure between the delivery pressure of the pump and the maximum load pressure among the plurality of hydraulic actuators becomes equal to the differential pressure target delivery amount Q ⁇ p.
  • the fourth device since the input limiting target delivery amount QT is not selected by the third device, the fourth device will not calculate the compensation value Qns, and the total consumable flow compensating control for restricting the flow rate through the flow control valve will not be performed.
  • the delivery amount of the pump is controlled while being limited such that it becomes equal to the input limiting target delivery amount QT.
  • the fourth device calculates the compensation value Qns, and the total consumable flow compensating control is performed for restricting the flow rate through the flow control valve.
  • the differential pressure target delivery amount Q ⁇ p and the input limiting target delivery amount QT are independently calculated as the target delivery amount Qo for the pump, and the total consumable flow compensating control is carried out only when the input limiting target delivery amount QT is selected. Therefore, the load-sensing control and the total consumable flow compensating control will not occur simultaneously. Specifically, in the condition where the delivery amount of the pump is less than its available maximum delivery amount (the input limiting target delivery amount QT), the load-sensing control is carried out, while in the condition where it reaches the available maximum delivery amount, the total consumable flow compensating control is carried out. This enables smooth increases or decreases in the flow rates supplied to the respective hydraulic actuators and hence improve the operability. It is also possible to prevent a hunting phenomenon due to interference between the load-sensing control and the total consumable flow compensating control, resulting in the stable control.
  • the consumable flow rate which is passed through the pressure compensated flow control valve to the associated hydraulic actuator is determined based on both a throttle opening command value for a flow control valve given by the operation signal from the control means, and a differential pressure command value across the flow control valve given to the pressure balance valve in the form of the compensation value Qns from the fourth device.
  • the operation signal modifying factor ⁇ is calculated from the compensation value Qns and the operation signal from the control device is modified using the operation signal modifying factor ⁇ to control the pressure compensated flow control valve
  • the above differential pressure command value is included in the throttle opening command value for the flow control valve given by the modified operation signal
  • the consumable flow rate is determined by the modified operation signal (throttle opening command value).
  • the new target delivery amount Qo is always calculated from the preceding target delivery amount Qo-1 and the transition is hence smoothed when the pump is shifted from the condition where it is controlled following the differential pressure target delivery amount Q ⁇ p to the condition where it is controlled following the input limiting target delivery amount QT, or vice versa.
  • the pump will not be subjected to rush operation at the time of shifting the control mode, and more stable control is ensured.
  • the fourth device calculates a deviation between the compensation value Qns and the preset offset value and outputs the resulting value Qnso as the final compensation value.
  • the total consumable flow rate determined by the pressure compensated flow control valve under control using Qnso becomes slightly greater than the available maximum delivery amount of the pump by an extent corresponding to the offset value, and hence there produces a corresponding free flow rate in the delivery amount of the pump, which can pass into the hydraulic actuator(s) on the lower pressure side.
  • most of the flow rate is under the total consumable flow compensating control, which ensures a function to certainly supply the hydraulic fluid to the actuator(s) on the higher pressure side as well, for achieving the combined operation.
  • the pressure compensated flow control valve is hydraulically controlled directly with the differential pressure between the delivery pressure of the pump and the maximum load pressure among the actuators, as mentioned above, the total consumable flow rate is coincident with the actually consumed total flow rate.
  • the pressure compensated flow control valve is controlled using a calculated value and hence the total consumable flow rate can be selected optionally. For example, as set forth above, it is possible to make a control system such that the total consumable flow rate becomes larger than the delivery amount of the pump. In this case, the total consumable flow rate can exceed the actually consumed total flow rate.
  • the present invention is applicable to not only such a mode, but also another mode in which the throttle openings of the respective pressure compensated flow control valves are reduced to be slightly different from each other.
  • FIG. 1 is a schematic view showing a control system for a hydraulic drive circuit according to one embodiment of the present invention, including the hydraulic drive circuit itself;
  • FIG. 2 is a sectional view showing the structure of a differential pressure gauge for the control system
  • FIG. 3 is a schematic view showing the configuration of a delivery amount control device in the control system
  • FIG. 4 is a sectional view showing the structure of a proportional solenoid valve in the control system
  • FIG. 5 is a schematic view showing the configuration of a control unit as a main component of the control system
  • FIG. 6 is a flowchart showing control programs used in the control unit
  • FIG. 7 is a graph showing an input torque limiting function used for determining an input limiting target value
  • FIG. 8 is a block diagram showing the procedure of determining a differential pressure target delivery amount from the differential pressure between the delivery pressure of a hydraulic pump and the maximum load pressure;
  • FIG. 9 is a block diagram showing the procedure of determining a total consumable flow compensating current from the target delivery amount deviation
  • FIG. 10 is a flowchart showing the procedure to control a delivery amount control based on both the delivery amount target value and the inclination angle signal;
  • FIG. 11 is a control block diagram showing the entire control procedure
  • FIG. 12 is a schematic view showing a control system according to a second embodiment of the present invention.
  • FIG. 13 is a graph showing an input torque limiting function used in the control system of FIG. 12;
  • FIG. 14 is a control block diagram of the control system of FIG. 12;
  • FIGS. 15A and 15B are a control block diagram of a control system for a hydraulic drive circuit according to a third embodiment of the present invention, including the hydraulic drive circuit;
  • FIG. 16 is a control block diagram of a control system for a hydraulic drive circuit according to a fourth embodiment of the present invention.
  • FIG. 17 is a control block diagram of a control system for a hydraulic drive circuit according to a fifth embodiment of the present invention.
  • FIG. 18 is a control block diagram of a control system for a hydraulic drive circuit according to a sixth embodiment of the present invention.
  • FIG. 19 is a control block diagram of a control system for a hydraulic drive circuit according to a seventh embodiment of the present invention.
  • FIG. 1 shows an overall arrangement of a load-sensing hydraulic drive circuit and a control system of the present invention.
  • the load-sensing hydraulic drive circuit comprises a variable displacement hydraulic pump 1 of the swash plate type, for example, first and second hydraulic actuators 2 and 3, driven by hydraulic fluid delivered from the hydraulic pump 1, a first flow control valve 4 and a first pressure balance valve 6 for pressure compensation both disposed between the pump 1 and the first actuator 2 to control the flow rate and direction of hydraulic fluid supplied to the first actuator 2 from the pump 1, and a second flow control valve 5 and a second pressure balance valve 7 for pressure compensation both disposed between the pump 1 and the second actuator 3 to control the flow rate and direction of hydraulic fluid supplied to the second actuator 3 from the pump 1.
  • a variable displacement hydraulic pump 1 of the swash plate type for example, first and second hydraulic actuators 2 and 3, driven by hydraulic fluid delivered from the hydraulic pump 1, a first flow control valve 4 and a first pressure balance valve 6 for pressure compensation both disposed between the pump 1 and the first actuator 2 to control the flow rate and
  • the first pressure balance valve 6 is connected at its inlet side to the pump 1 through a hydraulic fluid supply line 20, and at its outlet side to the flow control valve 4 through a line with a check valve 22.
  • the flow control valve 4 is connected at its inlet side to the pressure balance valve 6 and also to a tank 10 through a return line 24, and at its outlet side to the first actuator 2 through main lines 25, 26.
  • the second pressure balance valve 7 is connected at its inlet side to the pump 1 through a line 21 and the hydraulic fluid supply line 20, and at its outlet side to the flow control valve 5 through a line with a check valve 23.
  • the flow control valve 5 is connected at its inlet side to the pressure balance valve 7 and also to the tank 10 through a return line 29, and at its outlet side to the second actuator 3 through main lines 27, 28.
  • the pressure balance valve 6 is of a pilot operated type having two closing-direction working pilot pressure chambers 6a, 6b and an opening-direction working pilot chamber 6c located in opposite relation.
  • the inlet pressure of the flow control valve 4 is applied to the closing-direction working-pilot pressure chamber 6a through a line 30, the outlet pressure of a proportional solenoid valve 9 (later described) is applied to the other pressure chamber 6b through a line 31, and the pressure (later described) between the flow control valve 4 and the first actuator 2 is applied to the opening-direction working pilot pressure chambers 6c through a line 32a.
  • the pressure balance valve 6 further includes a spring 6d for urging the valve 6 in the opening direction.
  • the pressure balance valve 7 is also constructed in a like manner. More specifically, the pressure balance valve 7 is of a pilot operated type having two closingdirection working pilot pressure chambers 7a, 7b and an opening-direction working pilot chamber 7c located in opposite relation.
  • the inlet pressure of the flow control valve 5 is applied to the closing-direction working pilot pressure chambers 7a, through a line 33
  • the outlet pressure of the proportional solenoid valve 9 is applied to the other pressure chamber 7b through a line 34
  • the pressure between the flow control valve 5 and the second actuator 3 is applied to the openingdirection working pilot pressure chambers 7c through a line 35a.
  • the pressure balance valve 7 further includes a spring 7d for urging the valve 7 in the opening direction.
  • the pressure balance valve 6 operates as follows. When the pressure of the proportional solenoid valve 9 is 0 (zero), the pressure balance valve 6 is subjected to the inlet pressure of the flow control valve 4 introduced to its pilot chamber 6a through the line 30, in one direction, and to the outlet pressure of the flow control valve 4 introduced to its pilot chamber 6c through the line 32a and the resilient urging force of the spring 6d, in the opposite direction. Therefore, the pressure balance valve 6 always controls the flow rate from the pump 1 so that the differential pressure between the inlet pressure and the outlet pressure of the flow control valve 4 is held a a constant value corresponding to the resilient urging force of the spring 6d.
  • the pressure balance valve 6 functions as a flow control valve for pressure compensation.
  • the pressure balance valve 7 also operates in a like manner.
  • the proportional solenoid valve 9 when the proportional solenoid valve 9 produces a pressure, this pressure is transmitted to the pressure balance valves 6, 7 through the lines 31, 34 and acts to counter the resilient urging forces of the opposing springs 6d, 7d. Stated otherwise, the pressure balance valves 6, 7 are each controlled so as to reduce the differential pressure between the inlet pressure and the outlet pressure of the flow control valves 4, 5 in proportion to a pressure rise in lines 31 and 34, and hence the flow rate through the flow control valves 4, 5 is reduced. Thus, controlling the pressure of the proportional solenoid valve 9 makes it possible to restrict the flow rates through the flow control valves 4, 5 and carry out total consumable flow compensating control thereof.
  • the flow control valves 4 and 5 are of a pilot operated type having opposed pilot chambers connected to pilot lines 36a, 36b and 37a, 37b, respectively, and are controlled with pilot pressures transmitted through pilot lines in response to operation signals from the respective control levers (not shown).
  • the flow control valve 4 and the pressure balance valve 6 jointly constitute a single pressure compensated flow control valve.
  • the operation signal from the associated control lever gives a throttle opening command value for the flow control valve 4, while the pressure applied to the pressure balance valve 6 from the proportional solenoid valve 9 and the setting value of the spring 6d give a command value for the differential pressure across the flow control valve 4.
  • the throttle opening command value and the differential pressure command value for the flow control valve 4 determine a consumable flow rate that is to be passed from the pressure compensated flow control valve 4 to the hydraulic actuator 2, and the throttle opening of the flow control valve and the throttle opening of the pressure balance valve are so controlled as to achieve the consumable flow rate.
  • the actual flow rate through the pressure compensated flow control valve that is, the consumed flow rate through the hydraulic actuator, is thus controlled.
  • the flow control valve 5 and the pressure balance valve 7 jointly constitute another pressure compensated flow control which operates in a like manner.
  • pilot lines 32, 35 are Also connected to the flow control valves 4, 5 for picking up the load pressures of the first and second actuators 2, 3, respectively.
  • the pilot lines 32, 35 are arranged such that they are connected in the interior of the flow control valves 4, 5 to the return lines 24, 29 in a neutral state and to the main lines of the actuators 2, 3 coupled to the pump 1 in an operated state.
  • the higher one of the pressures in the lines 32, 35 is selected by a higher-pressure selector valve 12 and then introduced to a differential pressure gauge 43 through a line 38. Further introduced to the differential pressure gauge 43 is the delivery pressure of the pump 1 through a line 39.
  • the differential pressure gauge 43 detects the differential pressure between the delivery pressure of the pump 1 and the higher load pressure (maximum load pressure), and then outputs a differential pressure signal ⁇ P.
  • the differential pressure gauge 43 has such a construction as shown in FIG. 2 by way of example.
  • the differential pressure gauge 43 includes a body 50 having hydraulic fluid supply ports 47, 48 connected to the lines 38, 39, respectively, and a hydraulic fluid discharge port 49 connected to the tank 10 through a line 41, a cylinder 51 fitted in the body 50, a piston 52 accommodated in the cylinder 51 and having two pressure receiving surfaces 52a, 52b of equal area which are opposite to each other and subjected to the different pressures from the supply ports 47, 48, respectively, a shaft 53 made of a non-magnetic substance and transmitting a displacement and force of the piston 52, a spring 54 accommodated in the cylinder 51 for receiving the force of the piston 52 and giving a displacement proportional to the received force to the piston 52, a case 55 made of a non-magnetic substance and fitted to the cylinder 51, a core 56 made of a magnetic substance, attached to the distal end of the shaft 53 and accommodated in the case 55 for being displaced in the case 55 through the same
  • the pump delivery pressure P and the maximum load pressure Pam act on the pressure receiving surfaces 52a, 52b of the piston 52 through the supply ports 47, 48, respectively.
  • the force of A ⁇ (P-Pam) acts on the piston 52 upward in the figure because of P>Pam. That force causes the piston 52 to be displaced against the springs 54, 60 which are in their pre-compressed state to resiliently support the piston 52, so does the core 56.
  • the springs 54, 60 have their spring constants K1, K2, the displacement S is expressed by:
  • the displacement sensor 57 converts the displacement to an electric signal, and the amplified signal is output from the amplifier 59.
  • the displacement sensor 57 is preferably of a contactless type such as a differential transformer type or magnetic resistor element type, for example, because of the presence of oil deposited around the core 56. For this reason, the shaft 53 and the case 55 are both made of a non-magnetic substance.
  • the displacement sensor of any such type has a linear relationship between the displacement S and an electric signal level E, i.e., a simple proportional relationship. Letting the proportional constant to be K, therefore, the electric signal level E is expressed by:
  • the electric signal level E has a value proportional to the differential pressure (P-Pam) between the pump delivery pressure and the maximum load pressure, thereby providing the differential pressure signal ⁇ P.
  • the two pressures on the opposite pressure receiving surfaces of the piston 52 produce the differential pressure therebetween, making it is possible to avoid errors caused by non-linearity of the output from the pressure sensor with respect to the pressure and hysteresis upon rise and fall of the pressure.
  • errors would result in the case where the respective pressures are introduced to separate pressure sensors to produce electric signals and the difference in level between those two electric signals is then obtained to produce an electric signal corresponding to the differential pressure. Consequently, the differential pressure can be measured with a high degree of accuracy even under condition of higher pressure.
  • the differential pressure gauge 43 is merely needed to measure the differential pressure only in case of P>Pam in the illustrated embodiment, the spring 60 may be dispensed with.
  • the structure is simplified and the relationship between the output electric signal level E and the differential pressure is expressed by:
  • a pressure detector 14 for detecting the delivery pressure of the pump 1 and producing an output pressure signal P.
  • the pump 1 is provided with an inclination angle gauge 15 which detects an inclination angle of the displacement volume varying mechanism such as a swash plate and outputs an inclination angle signal Q ⁇ .
  • the pump 1 is controlled substantially constant in the rotational speed thereof, and thus the inclination angle signal Q ⁇ indicates the delivery amount of the pump 1.
  • the delivery amount of the pump 1 is controlled by a delivery amount controller 16 which is coupled to the displacement volume varying mechanism.
  • the delivery amount controller 16 can be constructed, for example, in the form of an electro-hydraulic servo-type hydraulic drive device as shown in FIG. 3.
  • the delivery amount controller 16 has a servo piston 16b which drives a displacement volume varying mechanism 16a, such as a swash plate, swash shaft or the like, of the variable displacement hydraulic pump 1, the servo piston 16b being accommodated in a servo cylinder 16c.
  • a cylinder chamber of the servo cylinder 16 is divided by a servo piston 16b into a left-hand chamber 16d and a righthand chamber 16e, and the lefthand chamber 16d is formed to have the cross-sectional area D larger than that d of the righthand chamber 16e.
  • Designated at 8 is the pilot pump or hydraulic source for supplying hydraulic fluid to the servo cylinder 16c.
  • the hydraulic source 8 and the lefthand chamber 16d of the servo cylinder 16c is intercommunicated through a line 16f
  • the hydraulic source 8 and the righthand chamber 16e of the servo cylinder 16c is intercommunicated through a line 16i.
  • These lines 16f and 16i are communicated to the tank 10 through a return line 16j.
  • a solenoid valve 16g is disposed in the line 16f intercommunicating the hydraulic source 8 and the lefthand chamber 16d of the servo cylinder 16c
  • another solenoid valve 16h is disposed in the return line 16j.
  • These solenoid valves 16g, 16h are normally-closed solenoid valves, automatically returning to a closed state when deenergized, and their state is switched by a load-sensing control signal Q'o from a control unit 40, described later.
  • the inclination angle of the displacement volume varying mechanism 16a of the pump 1 is held constant and hence the delivery amount thereof is also held constant.
  • the solenoid valve 16h is energized (turned on) for being brought into a switched position B
  • the lefthand chamber 16d of the servo cylinder 16c is communicated with the tank 10, so that the servo piston 16b is moved leftward in FIG. 3 under the action of the pressure in the righthand chamber 16e upon reduction of the pressure in the lefthand chamber 16d.
  • the inclination angle signal Q ⁇ output from the inclination angle gauge 15 is controlled to have a level corresponding to a target delivery amount Qo calculated by the control unit 40, as described later.
  • the proportional solenoid valve 9 can be constructed, for example, as shown in FIG. 4.
  • the illustrated proportional solenoid valve 9 contains by a proportional solenoid pressure-reducing valve, and includes a proportional solenoid part 62 and a pressure-reducing valve part 63.
  • the solenoid part 62 has a known structure comprising a solenoid with terminals 64a, 64b, and an iron core.
  • the input to terminals 64a, 64b is a total consumable flow compensating control signal Qns, described later, from the control unit 40.
  • the pressure-reducing valve 63 includes a body 71 having a hydraulic supply port 67 connected to an auxiliary pump 8 through a supply line 66, a hydraulic fluid discharge port 69 connected to the tank 10 through a return line 68, and a hydraulic outlet port 70 connected to the pilot lines 31, 34.
  • a spool 72 disposed in the body 71, having end faces 72a, 72b opposite to each other and formed with an internal passage 72c, and a push rod 73 engaging at one end with the iron core of the proportional solenoid part 62 and abutting at the other end against the end face 72a of the spool 72.
  • the spool 72 moves leftward to communicate the internal passage 72c with the discharge port 69, so that the outlet port 70 and the discharge port 69 are communicated with each other through the internal passage 72c.
  • the hydraulic pressure in the outlet port 70 is reduced and the force acting on the end face 72b of the spool 72 is also reduced.
  • the spool 72 is moved rightward again in the figure.
  • the pressure in the supply line 66 is designed to always stand at a constant level set by a relief valve 11.
  • the pressure signal P from the pressure detector 14, the inclination angle signal Q ⁇ from the inclination angle gauge 15, and the differential pressure signal ⁇ P from the differential pressure gauge 43 are input to the control unit 40 which generates the total consumable flow compensating control signal Qns and the load-sensing control signal Q'o, and then outputs them to the proportional solenoid valve 9 and the delivery amount controller 16, respectively.
  • the control unit 40 comprises a microcomputer and includes, as shown in FIG. 5, an A/D converter 40a for converting the pressure signal P output from the pressure detector 14, the inclination angle signal Q ⁇ output from the inclination angle gauge 15, and the differential pressure signal ⁇ P output from the differential pressure gauge 43 to respective digital signals.
  • Control unit 40 also comprises a central processing unit 40b, a memory 40c for storing a program for the control procedure, a D/A converter 40d for outputting analog signals, an I/O interface 40e for outputting signals, an amplifier 40f connected to the proportional solenoid valve 9, and amplifiers 40g, 40h connected to the solenoid valves 16g, 16h, respectively.
  • the control unit 40 calculates a delivery amount target value Qo for the variable displacement hydraulic pump 1 based on the control program stored in the memory 40c, and then outputs the loadsensing control command signal Q'o from the amplifiers 40g, 40h to the solenoid valves 16g, 16h of the delivery amount control 16, respectively, through the I/O interface 40e.
  • the position of the servo piston 3 is controlled by on-off servo control using an electrohydraulic servo technique so that the inclination angle signal Q ⁇ has a level corresponding to the delivery amount target value Qo, as explained above.
  • the control unit 40 also calculates a total consumable flow compensating value based on a control program stored in the memory 40c, and outputs the control command signal Qns from the amplifier 40f to the solenoid proportional control valve 9 through the D/A converter 40d. This causes the proportional solenoid valve 9 to produce a pressure in proportion to the command signal Qns, as explained above.
  • the control unit 40 reads and stores therein, as conditions of the hydraulic drive system, the delivery pressure P of the pump 1, the inclination amount Q ⁇ of the pump 1, and the differential pressure ⁇ P between the maximum load pressure Pam and the delivery pressure P from the outputs of the pressure detector 14, the inclination angle gauge 15 and the differential pressure gauge 43, respectively.
  • an input limiting target delivery amount QT is determined based on both the output pressure P of the pressure detector 14 and an input torque limiting function f(P) previously input in the memory.
  • FIG. 7 shows the input torque limiting function.
  • the X-axis represents the output pressure P and the Y-axis represents the input limiting target delivery amount QT based on the input torque limiting function f(P).
  • the input torque of the pump 1 is in proportion to the product of the delivery pressure P and the inclination amount Q ⁇ of the pump 1. Accordingly, the input torque limiting function f(P) is given by a hyperbolic curve or an approximate hyperbolic curve.
  • f(P) is such a function as expressed by the following equation:
  • the input limiting target delivery amount QT can be determined.
  • FIG. 8 is a block diagram showing a method of determining the differential pressure target delivery amount Q ⁇ p from the differential pressure signal ⁇ P of the differential pressure gauge 43.
  • the differential pressure target delivery amount Q ⁇ p is determined based on the following equation: ##EQU1## where KI: integration gain
  • this example calculates the differential pressure target delivery amount Q ⁇ p using an integration control technique applied to a deviation between the target differential value ⁇ Po and the actual difference pressure.
  • a block 120 calculates KI( ⁇ Po- ⁇ P) from the differential pressure ⁇ P for determining an increment ⁇ Q ⁇ p of the differential pressure target delivery amount per one unit of control cycle time, and a block 121 obtains the equation (2) by adding the above ⁇ Q ⁇ p and the delivery amount target value Qo-1 in the preceding control cycle.
  • Q ⁇ p has been determined using the integral control technique applied to ⁇ Po- ⁇ P in the foregoing embodiment, it may be determined using any other suitable technique.
  • the proportional control technique expressed by;
  • Kp is a proportional gain or a proportional plus integral control technique can be performed by using the sum of the equations (2) and (3).
  • the differential pressure target delivery amount Q ⁇ p is determined in step 102.
  • step 104 determines whether the deviation ⁇ Q is positive or negative. If the deviation ⁇ Q is positive, the process goes to step 105 to select QT as the delivery amount target value Qo. If the deviation ⁇ Q is negative, it goes to step 106 to select Q ⁇ p as the delivery amount target value Qo. In other words, the lesser of the differential pressure target delivery amount Q ⁇ p and the input limiting target delivery amount QT is selected as the delivery amount target value Qo, so that the delivery amount target value Qo will not exceed the input limiting target delivery amount QT determined by the input torque limiting function f(P).
  • FIG. 9 is a block diagram showing a method to calculate the compensation value Qns from the target delivery amount deviation ⁇ Q.
  • an compensation value Qns is determined using the integral control technique based on the following equation: ##EQU2## where KIns: integral gain
  • Qns-1 total consumable flow compensation value Qns output in the preceding control cycle
  • the compensation value increment ⁇ Qns per one unit of control cycle time i.e., KIns ⁇ Q
  • KIns ⁇ Q the compensation value increment ⁇ Qns per one unit of control cycle time
  • the increment is then added in an adder 131 to the compensation value Qns-1 output in the preceding control cycle, thereby to determine an intermediate value Q'ns.
  • Qnsmax and Q'nsc are values determined by the maximum inclination angle of swash plate of the pump 1, i.e., the maximum delivery amount thereof.
  • the compensation value Qns has been determined using an integral control technique in the foregoing embodiment, the relationship between Qns and ⁇ Q may be determined using a proportional control technique or the proportional plus integral control technique, as with the above case of the differential pressure target delivery amount Q ⁇ p.
  • the control unit 40 creates the command signal Q'o for the delivery amount control 16 based on the delivery amount target value Qo of pump 1 and the inclination angle signal Q ⁇ output from the inclination angle gauge 15 which are obtained in steps 105, 106, respectively.
  • the command signal Q'o is output to the delivery amount controller 16 through the I/O interface 40e and the amplifiers 40g, 40h of the control unit 40, as shown in FIG. 5, so that the inclination amount Q ⁇ of the pump 1 becomes equal to the delivery amount target value Qo.
  • FIG. 10 shows a flowchart of the control process carried out in step 108.
  • step 141 determines whether an absolute value of the deviation Z is larger or smaller than a value ⁇ preset for specifying the dead zone. If the absolute value of the deviation Z is larger than the preset value ⁇ , the process flow goes to step 142 to determine whether the deviation Z is positive or negative. If the deviation Z is positive, it goes to step 143 for outputting the command signal Q'o which turns ON the solenoid valve 16g of the delivery amount control 16 and turns OFF the solenoid valve 16h thereof.
  • the inclination angle of the pump 1 is increased so that the inclination angle signal Q ⁇ is controlled to be coincide with the target command signal Qo.
  • the process flow goes to step 144 for outputting the command signal Q'o which turns OFF the solenoid valve 16g and turns ON the solenoid valve 16h. This reduces the inclination angle of pump 1, so that the inclination angle signal Q ⁇ is controlled to be coincide with the target command signal Qo.
  • the process flow goes to step 145 where the solenoid valves 16g and 16h are both turned OFF. This causes the inclination angle of pump 1 to stand constant.
  • the differential pressure target delivery amount Q ⁇ p is selected as a delivery amount target value Qo in step 106 if the differential pressure target delivery amount Q ⁇ p is smaller than the input limiting target delivery amount QT, the delivery amount of the pump 1 is controlled to be equal to the differential pressure target delivery amount Q ⁇ p, and the differential pressure between the delivery pressure of the pump 1 and the maximum load pressure out of the plural actuators 2, 3 which is held constant.
  • the load-sensing control is effected.
  • the input limiting target delivery amount QT is selected as a delivery amount target value Qo in the step 105, and therefore the delivery amount of the pump is so controlled as not to exceed the input limiting target delivery amount QT.
  • the delivery amount of the pump is subjected to input limiting control.
  • step 109 an output current to the proportional solenoid valve 9 through the D/A converter 40d and the amplifier 40f of the control unit 40, as shown in FIG. 5, is controlled to be equal to Qns for controlling the pressure balance valves 6, 7 shown in FIG. 1.
  • the target current Qns is set 0 in block 132 (FIG. 9) in step 107.
  • the target current Qns is increased with an increase of the target delivery amount deviation ⁇ Q until the maximum value of Qnsmax in step 107, so that the throttle openings of the pressure balance valves 6, 7 are restricted in response to increase of the target delivery amount deviation ⁇ Q.
  • the total consumable flow compensating control is effected.
  • a block 200 corresponds to step 101 in FIG. 6 in that it calculates the input limiting target delivery amount QT based on the input torque limiting function shown in FIG. 7.
  • Blocks 201, 202, 203 correspond to step 102.
  • the addition block 201 and the proportional calculation block 202 correspond to the differential pressure target delivery amount increment calculation block 120 in FIG. 8, and the addition block 203 corresponds to the adder 121 in FIG. 8.
  • the differential pressure target value Q ⁇ p is calculated through these three blocks.
  • Block 204 corresponds to steps 104, 105, and 106 in FIG. 6 in that it selects the lesser of the two target delivery amounts QT and Q ⁇ p as the delivery amount target value Qo.
  • Blocks 205, 206, 207, 208 correspond to step 107 in FIG. 6.
  • the addition block 205 and the proportional calculation block 206 correspond to the total consumable flow compensation value increment calculation block 131 in FIG. 9, respectively
  • the addition block 207 corresponds to the limiter 132 in FIG. 9.
  • the total consumable flow compensation value Qns is calculated through those three blocks.
  • Blocks 209, 210, 211 correspond to step 108 in FIG. 6.
  • the addition block 209 corresponds to the step 140 in FIG. 10
  • the blocks 210 and 211 correspond to the steps 141-145 in FIG. 10 in outputting the command signals Q'o to the respective solenoid valves 16g, 16h.
  • the input limiting target delivery amount QT and the differential pressure target delivery amount Q ⁇ p are calculated independently of each other as the target delivery amount Qo of pump 1, and only if the differential pressure target delivery amount Q ⁇ p exceeds the input limiting target delivery amount QT, the total consumable flow compensating control is carried out. Therefore, when the differential pressure target delivery amount is smaller than the input limiting target delivery amount and hence there is no need of total consumable flow compensating control, the total consumable flow compensating control will not be carried out even if the differential pressure ⁇ P is reduced due to a response lag in the delivery amount control 16 for the pump 1. Therefore, the throttle openings of the pressure balance valves 6, 7 will not be restricted.
  • the flow control valves 4, 5 can provide the flow rates as exactly specified by the associated control levers. Further, the load-sensing control and the total consumable flow compensating control are not effected concurrently, and this prevents a hunting phenomenon from occurring due to interference therebetween, and hence ensures stable control of the hydraulic actuators 2, 3.
  • QT has been determined from the delivery pressure P and the input torque limiting function f(P).
  • FIGS. 12 and 13 show such an embodiment in which the identical members to those in FIG. 1 are designated with the same reference numerals.
  • an internal combustion engine 150 for driving a plurality of pumps including a hydraulic pump 1 is shown.
  • Fuel is supplied to engine 150 by a fuel injection pump 151.
  • the target speed for engine 150 is set by an accelerator 152.
  • the engine 150 has a speed sensor 153 on its output shaft which detecting rotational speed .
  • a target engine speed signal Nr from accelerator 152 and an actual engine speed signal Ne from the speed sensor 153 are input to a control unit 154 for the engine 150 for determining an engine speed deviation ⁇ N therebetween.
  • Also input to the control unit 154 is a rack displacement signal from a rack displacement detector 155 for the fuel injection pump 151.
  • the control unit 154 Based on the engine speed deviation ⁇ N and the rack displacement signal, the control unit 154 calculates a target rack displacement for the fuel injection pump 151 and then outputs a rack operating signal to the fuel injection pump 151. Further, the control unit 154 outputs the engine speed deviation ⁇ N to the control unit 40 for the hydraulic pump 1 as well.
  • the control unit 40 stores therein, as the input limiting function for the pump 1, an input torque limiting function f1(P, ⁇ N) with parameters of the delivery pressure P of the pump 1 and the engine speed deviation ⁇ N of the internal combustion engine 150.
  • FIG. 13 shows the input torque limiting function f1(P, ⁇ N).
  • the input torque limiting function f1(P, ⁇ N) reduces the product of the target delivery amount QT and the delivery pressure P as the engine speed deviation ⁇ N is increased, thereby controlling the target delivery amount QT.
  • the input limiting target delivery amount QT is determined based on the engine speed deviation ⁇ N, the delivery pressure P and the input torque limiting function f1(P, ⁇ N). By so doing, the torque of pump 1 can be reduced with the increasing engine speed deviation ⁇ N.
  • a control block diagram of this embodiment is shown in FIG. 14.
  • block 250 compares the actual engine speed signal Ne from the speed sensor 153 with the target engine speed signal Nr from the accelerator 152 to calculate the engine speed deviation ⁇ N.
  • a block 251 is an input limiting target delivery amount calculation block which inputs the delivery pressure P and the engine speed deviation ⁇ N for calculating the input limiting target delivery amount QT from the input torque limiting function shown in FIG. 13. Other blocks are the same as those in FIG. 11.
  • the input torque limiting control of pump 1 is performed such that the product of the target delivery amount QT and the delivery pressure P is made smaller with the increasing engine speed deviation ⁇ N. It is thus possible to effectively utilize the output horsepower of the engine 150 at maximum.
  • FIGS. 15A and 15B A third embodiment of the present invention will be described with reference to FIGS. 15A and 15B.
  • the components similar to those in FIGS. 1 and 11 are denoted at the same reference numerals.
  • the flow control valve rather than the pressure balance valve, is controlled directly based on the total consumable flow compensation value Qns.
  • the pressure balance valves 6, 7 of the respective pressure compensated flow control valves are controlled using the compensation value Qns.
  • the consumable flow rates transmitted to the hydraulic actuators 2, 3 through the respective pressure compensated flow control valves are determined based on both the throttle opening command values for the flow control valves 4, 5 given by the operation signal from the associated control levers, and the differential pressure command values across the flow control valves given to the pressure balance valves 6, 7 as the compensation values Qns.
  • the operation signals of the control levers are modified using the compensation value Qns to include the differential pressure command values into the respective throttle opening command values for the flow control valves 6, 7, whereby the consumable flow rates are determined by the resulting throttle opening command values.
  • control levers which output operation signals Qa1, Qa2 of the hydraulic actuators 2, 3 when operated, respectively.
  • a control unit 40A serves, in addition to the function of the control unit 40 in FIG. 1, to input the operation signals Qa1, Qa2 from the control levers 70, 71, convert the input signals to drive signals Q'a1+, Q'a1- and Q'a2+, Q'a2- for proportional solenoid valves 9a-9d, and then output them, respectively.
  • the proportional solenoid valves 9a-9d produce pilot pressures for operating the flow control valves 4, 5 proportional to the drive signals Q'a1+, Q'a1-, Q'a2+, Q'a2- output from the control unit 40A.
  • the opening directions and degrees of opening O of flow control valves 4, 5 are controlled opening directions and degrees thereof with the pilot pressures output from the proportional solenoid valves 9a-9d. For example, when the drive signal Q'a1+ is output to the flow control valve 4, the flow control valve 4 is switched to the righthand side as shown with the pilot pressure output from the proportional solenoid valve 9a to take the throttle opening in proportion to Q'a1+. Similarly, when the drive signal Q'a1- is output, the flow control valve 4 is switched to the lefthand side as shown.
  • the pressure balance valves 6A, 7A are adjusted in their throttle openings to make the differential pressures between inlets and outlets of the flow control valves 4, 5 equal to values set by springs 6d, 7d, respectively.
  • the flow rates specified by the drive signals Q'a1- to Q'a2- are supplied to the actuators 2, 3.
  • control unit 40A the control procedure carried out in control unit 40A is represented in a control block diagram similar to FIG. 11.
  • the steps for the load-sensing control, up to calculation of Qns in the total consumable flow compensating control, are the same as those for control unit 40 in FIG. 11. Operation of control unit 40A will be described below by referring to the remaining part of the control block diagram.
  • control unit 40A determines an operation signal modifying factor ⁇ from Qns.
  • the relationship between the factor ⁇ and Qns is, for example, such that ⁇ is 1 near around 0 of Qns and then decreases as Qns increases, as shown in block 400. Note that the minimum value of ⁇ should be larger than 0.
  • the operation signals Qa1, Qa2 from the control levers 70, 72 which have been input through the A/D converter 40a (see FIG. 5), are multiplied by the operation signal modifying factor ⁇ in multipliers 401a, 401b for generating the modified operation signals Qa1', Qa2', respectively.
  • the modified operation signals Q'a1-, Q'a2- are separated into respective ⁇ pairs by limiters 402a-402d to generate the proportional solenoid drive signals Q'a1+, Q'a1-, Q'a2+, Q'a2 which are output to the proportional solenoid valves 9a-9d.
  • the compensation value Qns is 0 and hence the operation signal modifying factor becomes 1. Therefore, the modified operation signals Q'a1, Q'a2 are coincident with the operation signals Qa1, Qa2 from the control levers 70, 71, and the flow control valves comes into the same conditions as the case where they are operated by the operation signals Qa1, Qa2.
  • saturation occurs if the total of flow rates demanded by the operation signals Qa1, Qa2 exceeds above the input limiting target delivery amount QT.
  • pump 1 is controlled with the input limiting target delivery amount QT.
  • the operation signal modifying factor ⁇ is made smaller as the compensation value Qns gradually increases from 0.
  • the operation signals Qa1, Qa2 are multiplied by the operation signal modifying factor ⁇ less than 1 in the multipliers 401a, 401b, so that the modified operation signals Q'a1, Q'a2 are gradually reduced.
  • the flow rates through the flow control valves 4, 5 are also reduced correspondingly.
  • the modifying factor ⁇ When the modifying factor ⁇ is reduced down to a level at which the total value of the modified operation signals Q'a1, Q'a2 coincides with the input limiting target delivery amount QT, the differential pressure signal ⁇ P is restored and the differential pressure target delivery amount Q ⁇ p is reduced to be coincident with the input limiting target delivery amount QT. Therefore, the target delivery amount deviation ⁇ Q becomes 0, whereupon an increase of the compensation value Qns and a reduction of the modifying factor ⁇ are brought into end.
  • operation signals from the control levers have been described as electric signals in the above embodiment, those operation signals may be replaced by hydraulic pilot signals and the hydraulic pressures of the pilot signals may be regulated through a proportional solenoid valve using the operation signal modifying factor ⁇ .
  • a fourth embodiment of the present invention will be described with reference to FIG. 16.
  • the delivery amount of the pump is controlled to deliver the input limiting target delivery amount QT to prevent interference between the load-sensing control and the total consumable flow compensating control.
  • the pump when the differential pressure target delivery amount Q ⁇ p is larger than the input limiting target delivery amount QT in the saturated condition, the pump is controlled to deliver the input limiting target delivery amount QT. Then, the flow rates through the flow control valves 4, 5 are controlled with the total consumable flow compensation value Qns corresponding to deficiency a in the demanded flow rates commanded by the operated amounts of the flow control valves 4, 5 as compared with the input limiting target delivery amount QT, whereby the saturated condition is solved.
  • the differential pressure target delivery amount Q ⁇ p is increased again above the input limiting target delivery amount QT, which in turn, increases the compensation value Qns, and hence reduces the flow rates through the flow control valves 4, 5. Then, the differential pressure target delivery amount Q ⁇ p is increased once again.
  • the above may occur repeatedly. In short, there is a possibility that the load-sensing control and the total consumable flow compensating control proceed simultaneously and interfere with each other, which leads to a hunting phenomenon.
  • FIG. 16 A control block diagram for a control unit 40B of this embodiment is shown in FIG. 16. In the figure, blocks of the same number as those in FIG. 11 carry out the same functions. Note that the component configuration in this embodiment is the same as that in FIG. 1.
  • a block 300 determines whether the total consumable flow compensating control is being performed or not, and then sets a total consumable flow compensating flag FQns. This decision is made based on the total consumable flow compensation value Qns, such that the total consumable flow compensating control is not being performed when Qns is equal to or less than 0, and is being performed when Qns is above 0.
  • the flag FQns is set to 1 or 0 dependent on whether or not the total consumable flow compensating control is being performed.
  • a block 204A is a minimum value selection block which determines which of the input limiting target delivery amount QT and the differential pressure target delivery amount Q ⁇ p is smaller, and then and outputs the smaller one as a delivery amount target value Qor.
  • Block 301 is a delivery amount target value selector switch for the pump. Upon receiving the total consumable flow compensating flag FQns, when FQns is 0 the switch selects the delivery amount target value Qor selected by the minimum value selection block 204A, and when FQns is 1 input limiting target delivery amount is selected to be QT. Then the selected value is outputted as a delivery amount target value Qo.
  • the differential pressure target delivery amount Q ⁇ p exceeds QT and hence the block 204A selects QT as the delivery amount target value Qor.
  • the target delivery amount deviation ⁇ Q becomes positive (+) and the compensation value Qns is increased.
  • the flag FQns is set to 1 and the delivery amount target value selector switch 301 selects the input limiting target delivery amount QT as the delivery amount target value Qo.
  • the pump 1 is controlled to the input limiting target delivery amount QT.
  • the flow rates through the flow control valves 4, 5 are reduced using the compensation value Qns which is coincident with the input limiting target delivery amount QT, with the result that the saturated condition is solved.
  • FIG. 16 operates in a like manner to that of FIG. 11.
  • the differential pressure target delivery amount Q ⁇ p is reduced and becomes smaller than the input limiting target delivery amount QT.
  • block 204A selects Q ⁇ p as the delivery amount target value Qor.
  • the target delivery amount deviation ⁇ Q becomes negative (-)
  • the total consumable flow compensation value Qns remains positive (+) and the flag FQns is held at 1 because Qns is gradually reduced in a transient range. Therefore, the delivery amount target value selector switch 301 selects the input limiting target delivery amount QT as the delivery amount target value Qo and the pump 1 is hence held controlled to QT.
  • the differential pressure target delivery amount Q ⁇ p becomes smaller than QT. But, the delivery amount target value Qo is held at QT because the flag FQns remains at 1 while the compensation value Qns assumes a positive (+) value. Therefore, Qns is gradually reduced while the delivery amount of the pump 1 is still held at QT, and this reduction continues until Qns becomes 0.
  • the delivery amount target value selector switch 301 selects the differential pressure target delivery amount Q ⁇ p as the delivery amount target value Qo. Thereafter, Q ⁇ p is controlled to be coincident with the total of demand flow rates commanded by the operation signals for the flow control valves 4, 5.
  • a fifth embodiment of the present invention will be described with reference to FIG. 17.
  • This embodiment is different from that of FIG. 16 in that the input limiting target delivery amount is calculated integrally rather than proportionally.
  • the component arrangement is, therefore, similar to that shown in FIG. 1 as with the embodiment of FIG. 16.
  • block 500 is a target delivery pressure calculation block which inputs the preceding delivery amount target value Qo-1 and calculates a currently allowable target delivery pressure Pr from the preset input limiting torque for the pump 1.
  • the target delivery pressure Pr is sent to a differential pressure calculation block 501 where the target delivery pressure Pr is compared with the current delivery pressure P to calculate a calculated differential pressure ⁇ P.
  • the differential pressure ⁇ P is multiplied by the integration gain K IP in an input limiting target delivery amount increment calculation block 502 to calculate an increment ⁇ Qps of the input limiting target delivery amount per one unit of control cycle time.
  • the increment ⁇ Qps of the input limiting target delivery amount and an increment ⁇ Q ⁇ p of the differential pressure target delivery amount are sent to a delivery amount increment minimum value selector block 204B that determines which of the two increments is smaller and then outputs the smaller one as a target delivery amount increment ⁇ Qor.
  • the delivery amount increment selector switch 301A selects the target delivery amount increment ⁇ Qor selected by the delivery amount increment minimum value selector block 204B when FQns is 0 and the input limiting target delivery amount increment ⁇ Qps when FQns is 1, and then outputs the selected one as a delivery amount increment ⁇ Qo.
  • the delivery amount increment ⁇ Qo selected by the delivery amount increment selector switch 301A is added in a block 503 to the delivery amount target value Qo-1 calculated in the preceding control cycle for calculating the delivery amount target value Qo in this cycle.
  • the input limiting target delivery amount increment ⁇ Qps and the differential pressure target delivery amount ⁇ Q ⁇ p are sent to a block 205A for calculating a signal indicative of the difference therebetween as the target delivery amount deviation ⁇ Q.
  • the flow through the blocks 201, 202, 204B, 301A, 503 are the same as that through the blocks 201, 202, 203, 204A, 301 in the load-sensing control of FIG. 16 for calculating the differential pressure target delivery amount.
  • the flow through the blocks 500, 501, 502, 204B, 301A, 503 is substituted for that through the blocks 200, 204A, 301 in FIG. 16 for calculating the input limiting target delivery amount.
  • both the embodiments of FIGS. 16 and 17 carry out the same function. Stated otherwise, in the load-sensing control of FIG. 17, the increment of the differential pressure target delivery amount calculated from control of the differential pressure is always compared with the increment of the input limiting target delivery amount calculated from the limiting torque, and the minimum value therebetween is added to the current pump delivery amount for determining how the pump delivery amount should be controlled based on which one of the differential pressure and the limiting torque is used.
  • the target delivery amount is also used in block 205A in FIG. 17 for calculating the target delivery amount deviation as with the block 205 in FIG. 16:
  • block 205A in FIG. 17 becomes equivalent to the block 205 in FIG. 16.
  • the remaining blocks subsequent to block 206 operates in the exactly same manner as those in FIG. 16.
  • This embodiment functions in a like manner to that of FIG. 16. Specifically, the total consumable flow compensation value Qns is determined based on the deviation ⁇ Q between the available delivery amount of the pump and the target delivery amount determined from the differential pressure, and the resulting Qns is employed to control the pressure balance valve for solving the saturated condition. Also, while the pressure balance value is under total consumable flow compensating control, the pump is controlled to the input limiting target delivery amount to avoid interference with the total consumable flow compensating control.
  • the new target delivery amount Qo is always calculated from the preceding target delivery amount Qo-1 and the transition is hence smoothed when the pump is shifted from the condition where it is controlled following the differential pressure target delivery amount to the condition where it is controlled following the input limiting target delivery amount, or vice versa. Accordingly, the pump will not be subject to any rush operation and can control more stably at the time of shifting the control mode.
  • FIG. 18 A sixth embodiment of the present invention will now be described with reference to FIG. 18.
  • the same components as those shown in FIG. 11 are denoted with the same reference numerals.
  • This embodiment is different from the foregoing ones in the arrangement of the block which calculates the total consumable flow compensation value Qns.
  • the output ⁇ P" of the half-wave rectifier 601 and the differential pressure deviation ⁇ P' are both input to a signal selector switch 602.
  • the signal selector switch 602 selects the value ⁇ P' when ⁇ Q is positive, i.e., in case of the differential pressure target delivery amount Q ⁇ P ⁇ the input limiting target delivery amount QT, and the value ⁇ P" when ⁇ Q is negative, i.e., in case of Q ⁇ p ⁇ QT, followed by outputting the selected one as an increment ⁇ Q'ns of an intermediate value.
  • This increment ⁇ Q'ns is added to the output Qns-1 of the preceding control cycle in the adder 207 to obtain the intermediate value Q'ns.
  • the value Q'ns is then sent to the limiter 208.
  • the limiter 208 prevents the value Q'ns from exceeding a maximum limit and outputs it as the total consumable flow compensation value Qns.
  • the signal selector switch 602 selects ⁇ P' (>0) as the intermediate value Q'ns and the pressure compensated flow control valve is controlled for compensation using the compensation value Qns produced from the positive ⁇ P'.
  • this embodiment can function similar to the first embodiment.
  • a seventh embodiment of the present invention will be described with reference to FIG. 19. Likewise, the same components in FIG. 19 as those shown in FIG. 11 are denoted at the same reference numerals. This embodiment is different from the foregoing ones in that the total consumable flow compensation value Qns is further modified.
  • a track apparatus of a hydraulic excavator for example, the hydraulic fluid is supplied to righthand and lefthand track motors through the associated pressure compensated flow control valves. But, the performance of this track apparatus would suffer if the foregoing total consumable flow compensating control is strictly performed. More specifically, when the hydraulic excavator is travelling straight, a slight difference in the supply amount of hydraulic fluid between the lefthand and righthand track motors occurs due to small variations in the individual components such as the pressure balance valves and the flow control valves. This makes rotational speeds of the track motors slightly different from each other, whereby the vehicle body will slowly turn to the right or left.
  • an adder 610 is provided in this embodiment to subtract a small offset value Qnsof from the compensation value Qns and the resulting difference is output as a final compensation value Qnso.
  • the total consumable flow rate given by Qnso becomes slightly greater than the available maximum delivery flow rate of the pump by an extent corresponding to the offset value Qnsof.
  • the system then produces a corresponding free flow rate in the delivery amount of the pump, which can pass into the track motor on the lower pressure side.
  • Such a free flow rate can be utilized advantageously depending on the situation. For example, if the vehicle body equipped with the above track apparatus tends to turn to the left slowly because of the fact that the righthand track motor is supplied with the larger supply flow rate than the lefthand track motor due to variations in the individual components, the righthand track motor would produce larger drive torque than the lefthand track motor.
  • the hydraulic pressure is increased on the righthand side which allows, the free flow rate caused by the offset value Qnsof to pass into the lefthand track motor under the lower load pressure.
  • the vehicle body is automatically released from its tendency to curve to the left and can travel straight.
  • this embodiment makes it possible to solve the drawback as would be experienced in case of strictly performing the total consumable flow compensating control.
  • the differential pressure target delivery amount Q ⁇ p and the input limiting target delivery amount QT are independently calculated as the target delivery amount Qo of the pump, and the total consumable flow compensating control is carried out only when the input limiting target delivery amount QT is selected. Therefore, where the delivery amount of the pump is less than its available maximum delivery amount (the input limiting target delivery amount QT), the load-sensing control is carried out, while in the condition where it reaches the available maximum delivery amount (the input limiting target delivery amount QT), the total consumable flow compensating control is carried out.
  • This enables a smooth increase or decrease the flow rates supplied to the respective hydraulic actuators and hence improves the operability. It is also possible to prevent a hunting phenomenon due to interference between the load-sensing control and the total consumable flow compensating control, resulting in stable control.
  • the new target delivery amount Qo is always calculated from the preceding target delivery amount Qo-1 and the transition is hence smoothed when the pump is shifted from the condition where it is controlled following the differential pressure target delivery amount Q ⁇ p to the condition where it is controlled following the input limiting target delivery amount QT, or vice versa, thereby ensuring more stable control.
  • the amount of consumable flow compensating control can be reduced.
US07/301,718 1988-01-27 1989-01-26 Control system for load-sensing hydraulic drive circuit Expired - Fee Related US4967557A (en)

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Cited By (70)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US5025625A (en) * 1988-11-10 1991-06-25 Hitachi Construction Machinery Co., Ltd. Commonly housed directional and pressure compensation valves for load sensing control system
US5056312A (en) * 1988-07-08 1991-10-15 Hitachi Construction Machinery Co., Ltd. Hydraulic drive system for construction machines
US5062350A (en) * 1989-03-22 1991-11-05 Hitachi Construction Machinery Co., Ltd. Hydraulic drive system for civil engineering and construction machine
US5077975A (en) * 1989-05-05 1992-01-07 Mannesmann Rexroth Gmbh Control for a load-dependently operating variable displacement pump
US5079919A (en) * 1989-03-30 1992-01-14 Hitachi Construction Machinery Co., Ltd. Hydraulic drive system for crawler mounted vehicle
US5083430A (en) * 1988-03-23 1992-01-28 Hitachi Construction Machinery Co., Ltd. Hydraulic driving apparatus
US5101629A (en) * 1989-02-20 1992-04-07 Hitachi Construction Machinery Co., Ltd. Hydraulic circuit system for working machine
US5101628A (en) * 1990-01-22 1992-04-07 Shin Caterpillar Mitsubishi Ltd. Energy regenerative circuit in a hydraulic apparatus
US5129230A (en) * 1990-06-19 1992-07-14 Hitachi Construction Machinery Co., Ltd. Control system for load sensing hydraulic drive circuit
US5138838A (en) * 1991-02-15 1992-08-18 Caterpillar Inc. Hydraulic circuit and control system therefor
US5150574A (en) * 1989-05-02 1992-09-29 Hitachi Construction Machinery Co., Ltd. Hydraulic drive system for civil engineering and construction machine
US5152140A (en) * 1989-10-11 1992-10-06 Hitachi Construction Machinery Co., Ltd. Pressure compensating valve spool positioned by difference in pressure receiving areas for load and inlet pressures
US5170625A (en) * 1989-07-27 1992-12-15 Hitachi Construction Machinery Co., Ltd. Control system for hydraulic pump
US5174114A (en) * 1990-02-28 1992-12-29 Hitachi Construction Machinery Co., Ltd. Hydraulic drive system for construction machine
US5177964A (en) * 1989-01-27 1993-01-12 Hitachi Construction Machinery Co., Ltd. Hydraulic drive traveling system
US5207059A (en) * 1992-01-15 1993-05-04 Caterpillar Inc. Hydraulic control system having poppet and spool type valves
US5209063A (en) * 1989-05-24 1993-05-11 Kabushiki Kaisha Komatsu Seisakusho Hydraulic circuit utilizing a compensator pressure selecting value
US5222417A (en) * 1991-01-22 1993-06-29 Fuji Jukogyo Kabushiki Kaisha Hydraulic control system of an automatic transmission for a motor vehicle
US5249421A (en) * 1992-01-13 1993-10-05 Caterpillar Inc. Hydraulic control apparatus with mode selection
US5267440A (en) * 1990-09-11 1993-12-07 Hitachi Construction Machinery Co., Ltd. Hydraulic control system for construction machine
US5267441A (en) * 1992-01-13 1993-12-07 Caterpillar Inc. Method and apparatus for limiting the power output of a hydraulic system
US5297381A (en) * 1990-12-15 1994-03-29 Barmag Ag Hydraulic system
US5307631A (en) * 1991-01-28 1994-05-03 Hitachi Construction Machinery Co., Ltd. Hydraulic control apparatus for hydraulic construction machine
US5351914A (en) * 1991-06-14 1994-10-04 Fuji Jukogyo Kabushiki Kaisha Hydraulic control system for aircraft
WO1994024407A1 (en) * 1993-04-19 1994-10-27 Bobbie Joe Bowden Automatic drilling system
US5394696A (en) * 1990-12-15 1995-03-07 Barmag Ag Hydraulic system
US5409038A (en) * 1991-01-23 1995-04-25 Kabushiki Kaisha Komatsu Seisakusho Hydraulic circuit including pressure compensating valve
US5442912A (en) * 1992-12-04 1995-08-22 Hitachi Construction Machinery Co., Ltd. Hydraulic recovery device
US5447027A (en) * 1993-03-23 1995-09-05 Hitachi Construction Machinery Co., Ltd. Hydraulic drive system for hydraulic working machines
US5493861A (en) * 1991-08-19 1996-02-27 Danfoss A/S Hydraulic system with pump and load
US5501136A (en) * 1993-06-24 1996-03-26 Voac Hydraulics Boras Ab Control system for a hydraulic motor
US5540049A (en) * 1995-08-01 1996-07-30 Caterpillar Inc. Control system and method for a hydraulic actuator with velocity and force modulation control
US5659485A (en) * 1993-07-02 1997-08-19 Samsung Heavy Industry Co., Ltd. Discharge flow control system and method in hydraulic pump
US5896930A (en) * 1997-01-27 1999-04-27 Kabushiki Kaisha Kobe Seiko Sho Control system in hydraulic construction machine
US5930996A (en) * 1997-10-02 1999-08-03 Hitachi Construction Machinery Co., Ltd. Auto-acceleration system for prime mover of hydraulic construction machine and control system for prime mover and hydraulic pump
US6209321B1 (en) * 1997-08-29 2001-04-03 Komatsu Ltd. Hydraulic controller for a working machine
US6282891B1 (en) * 1999-10-19 2001-09-04 Caterpillar Inc. Method and system for controlling fluid flow in an electrohydraulic system having multiple hydraulic circuits
US6408622B1 (en) * 1998-12-28 2002-06-25 Hitachi Construction Machinery Co., Ltd. Hydraulic drive device
US6516932B2 (en) * 2000-09-29 2003-02-11 New Holland North America, Inc. Electro-hydraulic clutch hysteresis compensation
US6526747B2 (en) * 2000-01-25 2003-03-04 Hitachi Construction Machinery Co., Ltd. Hydraulic driving device
US6662705B2 (en) 2001-12-10 2003-12-16 Caterpillar Inc Electro-hydraulic valve control system and method
KR100415781B1 (ko) * 2000-10-06 2004-01-24 에스엠시 가부시키가이샤 자기센서가 부착된 전환밸브
US6684636B2 (en) * 2001-10-26 2004-02-03 Caterpillar Inc Electro-hydraulic pump control system
US6701822B2 (en) 2001-10-12 2004-03-09 Caterpillar Inc Independent and regenerative mode fluid control system
US6715403B2 (en) 2001-10-12 2004-04-06 Caterpillar Inc Independent and regenerative mode fluid control system
US20040261407A1 (en) * 2003-06-30 2004-12-30 Hongliu Du Method and apparatus for controlling a hydraulic motor
US20050204735A1 (en) * 2004-03-17 2005-09-22 Kobelco Construction Machinery Co., Ltd. Hydraulic control system for working machine
US20060254671A1 (en) * 2002-12-09 2006-11-16 Endress + Hauser Flowtec Ag Method for filling a defined quantity of a medium into a container
US20070007039A1 (en) * 2002-11-05 2007-01-11 Sandvik Tamrock Oy Arrangement for controlling rock drilling
US20090126361A1 (en) * 2005-11-25 2009-05-21 Hitachi Construction Machinery Co., Ltd Pump Torque Controller of Hydraulic Working Machine
US20100218493A1 (en) * 2006-12-07 2010-09-02 Kazunori Nakamura Torque control apparatus for construction machine three-pump system
US20110020146A1 (en) * 2008-03-31 2011-01-27 Teruo Akiyama Rotation drive controlling system for construction machine
KR20110073711A (ko) * 2009-12-24 2011-06-30 두산인프라코어 주식회사 건설기계의 동력제어장치
US8056950B2 (en) 2004-09-24 2011-11-15 Stryker Corporation In-ambulance cot shut-off device
CN102607876A (zh) * 2012-04-13 2012-07-25 山东大学 适用于模型试验的多路高精度液压加卸载伺服控制系统
US20120251332A1 (en) * 2009-12-24 2012-10-04 Doosan Infracore Co., Ltd. Power control apparatus and power control method of construction machine
US20130121852A1 (en) * 2010-07-19 2013-05-16 Volvo Construction Equipment Ab System for controlling hydraulic pump in construction machine
US20130238178A1 (en) * 2012-03-07 2013-09-12 Clark Equipment Company Power management for a drive system
USRE44884E1 (en) 2004-09-24 2014-05-13 Stryker Corporation Ambulance cot with pinch safety feature
US20140299197A1 (en) * 2009-12-10 2014-10-09 Hydraforce, Inc. Method of controlling proportional motion control valve
US20140366955A1 (en) * 2013-06-13 2014-12-18 Caterpillar Global Mining America Llc Remote regulator for roof bolter
US9145905B2 (en) 2013-03-15 2015-09-29 Oshkosh Corporation Independent load sensing for a vehicle hydraulic system
US9518655B2 (en) 2013-01-29 2016-12-13 Deere & Company Continuously adjustable control management for a hydraulic track system
CN107355437A (zh) * 2017-06-28 2017-11-17 安徽柳工起重机有限公司 负载敏感回转缓冲阀及起重机回转液压系统
US10196131B2 (en) * 2016-02-16 2019-02-05 The Boeing Company Hydraulic system and method for an aircraft flight control system
US10989231B2 (en) * 2018-02-12 2021-04-27 Hawe Hydraulik Se Hydraulic valve assembly with forced circuit
US20210229924A1 (en) * 2018-07-06 2021-07-29 Sandvik Mining And Construction G.M.B.H. Hydraulic drive motor control system
US11220804B2 (en) * 2019-07-26 2022-01-11 Robert Bosch Gmbh Hydraulic pressurizing medium supply assembly for a mobile work machine, and method
US11608615B1 (en) * 2021-10-26 2023-03-21 Cnh Industrial America Llc System and method for controlling hydraulic valve operation within a work vehicle
US11834811B2 (en) 2021-10-25 2023-12-05 Cnh Industrial America Llc System and method for controlling hydraulic pump operation within a work vehicle

Families Citing this family (42)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPH02107802A (ja) * 1988-08-31 1990-04-19 Hitachi Constr Mach Co Ltd 油圧駆動装置
EP0432266B2 (en) * 1989-01-18 1997-08-13 Hitachi Construction Machinery Co., Ltd. Hydraulic driving unit for construction machinery
DE4005655C2 (de) * 1990-02-22 1993-10-07 Hydromatik Gmbh Vorrichtung zum Verstellen des Verdrängervolumens einer hydrostatischen Maschine
DE4005967C2 (de) * 1990-02-26 1996-05-09 Rexroth Mannesmann Gmbh Steueranordnung für mehrere hydraulische Verbraucher
GB2250611B (en) * 1990-11-24 1995-05-17 Samsung Heavy Ind System for automatically controlling quantity of hydraulic fluid of an excavator
FR2672944A1 (fr) * 1991-02-15 1992-08-21 Bennes Marrel Distributeur proportionnel et ensemble de commande d'une pluralite de recepteurs hydrauliques comportant pour chaque recepteur un tel distributeur.
DE4122164C1 (ko) * 1991-07-04 1993-01-14 Danfoss A/S, Nordborg, Dk
DE4235707B4 (de) * 1992-10-22 2007-10-18 Linde Material Handling Gmbh Hydrostatisches Antriebssystem
DE4235709A1 (de) * 1992-10-22 1994-04-28 Linde Ag Hydrostatisches Antriebssystem
JP2567193B2 (ja) * 1993-01-19 1996-12-25 三星重工業株式會社 油圧ポンプの吐出流量制御装置
JPH0742705A (ja) * 1993-07-30 1995-02-10 Yutani Heavy Ind Ltd 作業機械の油圧装置
KR950019129A (ko) * 1993-12-30 1995-07-22 김무 유압식 건설기계의 엔진-펌프 제어장치 및 방법
EP0795690B1 (en) * 1995-07-10 2001-12-05 Hitachi Construction Machinery Co., Ltd. Hydraulic driving device
JP3606976B2 (ja) * 1995-12-26 2005-01-05 日立建機株式会社 油圧作業機の油圧制御システム
JP3567051B2 (ja) * 1996-06-12 2004-09-15 新キャタピラー三菱株式会社 油圧アクチュエータ用の操作制御装置
JPH10196606A (ja) * 1996-12-27 1998-07-31 Shin Caterpillar Mitsubishi Ltd 油圧ポンプの制御装置
DE19745489B4 (de) * 1997-10-15 2004-07-22 O & K Orenstein & Koppel Aktiengesellschaft System zur lastdruckunabhängigen Steuerung und Lasthaltung mehrerer rotatorischer und/oder translatorischer Verbraucher
JP3854027B2 (ja) 2000-01-12 2006-12-06 日立建機株式会社 油圧駆動装置
DE10308289B4 (de) * 2003-02-26 2010-11-25 Bosch Rexroth Aktiengesellschaft LS-Wegeventilblock
DE10327519A1 (de) * 2003-06-17 2005-01-20 Ortlinghaus-Werke Gmbh Hydraulische Schaltung
CN100410549C (zh) * 2004-12-28 2008-08-13 东芝机械株式会社 液压控制装置
US8621855B2 (en) 2007-06-08 2014-01-07 Deere & Company Electro-hydraulic auxiliary mode control
KR101648982B1 (ko) * 2009-12-24 2016-08-18 두산인프라코어 주식회사 건설기계의 유압펌프 제어장치 및 제어방법
EP2686561A1 (en) * 2011-03-17 2014-01-22 Parker-Hannificn Corporation Electro-hydraulic system for controlling multiple functions
CN102734276B (zh) * 2012-06-28 2015-07-01 三一汽车起重机械有限公司 负载敏感电比例液压控制系统和工程机械
WO2014061507A1 (ja) * 2012-10-17 2014-04-24 日立建機株式会社 建設機械の油圧駆動装置
CN102878144A (zh) * 2012-10-26 2013-01-16 北京机械设备研究所 一种多通道负载敏感的液压控制回路
JP6005176B2 (ja) * 2012-11-27 2016-10-12 日立建機株式会社 電動式油圧作業機械の油圧駆動装置
CN103016466B (zh) * 2012-12-24 2015-03-25 中联重科股份有限公司 液压供油单元、液压泵站及液压供油单元的供油控制方法
CN104514765B (zh) * 2013-09-26 2018-02-23 陕西中大机械集团有限责任公司 用于控制摊铺机液压辅助系统的多路阀液压系统
EP2889433B1 (en) 2013-12-20 2019-05-01 Doosan Infracore Co., Ltd. System and method of controlling vehicle of construction equipment
KR102181125B1 (ko) * 2013-12-20 2020-11-20 두산인프라코어 주식회사 건설기계의 차량 제어 시스템 및 방법
CN105221277B (zh) * 2014-07-01 2017-11-21 徐工集团工程机械股份有限公司 起重机发动机功率控制方法及系统
CN105987033B (zh) * 2015-02-11 2018-08-14 徐工集团工程机械股份有限公司 低能耗多执行机构液压系统及挖掘机
JP7130474B2 (ja) * 2018-07-11 2022-09-05 住友建機株式会社 ショベル
JP7049213B2 (ja) * 2018-08-10 2022-04-06 川崎重工業株式会社 建設機械の油圧回路
CN110005650A (zh) * 2019-04-30 2019-07-12 郑钧译 一种双伺服浮动式自动液压机
US11225775B2 (en) * 2019-06-26 2022-01-18 Deere & Company Cycle time calibration
CN111765137B (zh) * 2020-06-12 2022-07-19 中联重科股份有限公司 复合动作检测模块、液压系统和起重机
SE545533C2 (en) * 2021-03-04 2023-10-17 Husqvarna Ab A hydraulic system for construction machines and a method for controlling the hydraulic system
CN113158344B (zh) * 2021-04-22 2024-02-23 上海三一重机股份有限公司 挖掘机辅助管路的控制方法、装置、作业机械及电子设备
CN114623112A (zh) * 2022-04-11 2022-06-14 华侨大学 一种新型纯电动抗流量饱和负载敏感系统及工程机械装置

Citations (14)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP0062072A1 (en) * 1980-10-09 1982-10-13 Hitachi Construction Machinery Co., Ltd. Method for controlling a hydraulic power system
US4395199A (en) * 1979-10-15 1983-07-26 Hitachi Construction Machinery Co., Ltd. Control method of a system of internal combustion engine and hydraulic pump
JPS58134342A (ja) * 1982-02-05 1983-08-10 Toshiba Corp 外部割込装置
JPS58174707A (ja) * 1982-04-06 1983-10-13 Daiden Kk 複数機器油圧駆動回路
DE3422165A1 (de) * 1983-06-14 1984-12-20 Linde Ag, 6200 Wiesbaden Hydraulische einrichtung mit einer pumpe und mindestens zwei von dieser beaufschlagten verbrauchern hydraulischer energie
JPS6011706A (ja) * 1983-06-14 1985-01-22 リンデ・アクチエンゲゼルシヤフト １つのポンプとこのポンプによつて負荷される少なくとも２つの液力作業装置とを有する液力式装置
EP0150308A2 (en) * 1984-01-30 1985-08-07 Trw Inc. Apparatus for controlling fluid flow
JPS60195339A (ja) * 1984-03-17 1985-10-03 Hitachi Constr Mach Co Ltd 油圧ポンプ駆動系制御装置
JPS60222601A (ja) * 1984-04-20 1985-11-07 Komatsu Ltd 油圧制御装置
JPS614848A (ja) * 1984-06-18 1986-01-10 Hitachi Constr Mach Co Ltd 原動機と油圧ポンプを含む系の制御装置
JPS6111429A (ja) * 1984-06-26 1986-01-18 Hitachi Constr Mach Co Ltd 原動機と油圧ポンプを含む系の制御装置
GB2171757A (en) * 1985-02-28 1986-09-03 Komatsu Mfg Co Ltd I c engine driven variable displacement pumping systems
US4747335A (en) * 1986-12-22 1988-05-31 Caterpillar Inc. Load sensing circuit of load compensated direction control valve
US4809504A (en) * 1986-01-11 1989-03-07 Hitachi Construction Machinery Co., Ltd. Control system for controlling input power to variable displacement hydraulic pumps of a hydraulic system

Family Cites Families (7)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4345435A (en) * 1980-05-05 1982-08-24 Sperry Corporation Power transmission
JPS6246724A (ja) * 1985-08-26 1987-02-28 Toyota Motor Corp 車両用自動クラツチの制御方法および装置
JPH0668281B2 (ja) * 1985-09-30 1994-08-31 株式会社小松製作所 流量制御方法及び装置
DE3546336A1 (de) * 1985-12-30 1987-07-02 Rexroth Mannesmann Gmbh Steueranordnung fuer mindestens zwei von mindestens einer pumpe gespeiste hydraulische verbraucher
DE3817218A1 (de) * 1987-06-11 1988-12-22 Mannesmann Ag Hydraulisches steuersystem fuer einen hydraulikbagger
AU603907B2 (en) * 1987-06-30 1990-11-29 Hitachi Construction Machinery Co. Ltd. Hydraulic drive system
JPH076521B2 (ja) * 1987-06-30 1995-01-30 日立建機株式会社 ロ−ドセンシング油圧駆動回路の制御装置

Patent Citations (15)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4395199A (en) * 1979-10-15 1983-07-26 Hitachi Construction Machinery Co., Ltd. Control method of a system of internal combustion engine and hydraulic pump
EP0062072A1 (en) * 1980-10-09 1982-10-13 Hitachi Construction Machinery Co., Ltd. Method for controlling a hydraulic power system
JPS58134342A (ja) * 1982-02-05 1983-08-10 Toshiba Corp 外部割込装置
JPS58174707A (ja) * 1982-04-06 1983-10-13 Daiden Kk 複数機器油圧駆動回路
US4617854A (en) * 1983-06-14 1986-10-21 Linde Aktiengesellschaft Multiple consumer hydraulic mechanisms
DE3422165A1 (de) * 1983-06-14 1984-12-20 Linde Ag, 6200 Wiesbaden Hydraulische einrichtung mit einer pumpe und mindestens zwei von dieser beaufschlagten verbrauchern hydraulischer energie
JPS6011706A (ja) * 1983-06-14 1985-01-22 リンデ・アクチエンゲゼルシヤフト １つのポンプとこのポンプによつて負荷される少なくとも２つの液力作業装置とを有する液力式装置
EP0150308A2 (en) * 1984-01-30 1985-08-07 Trw Inc. Apparatus for controlling fluid flow
JPS60195339A (ja) * 1984-03-17 1985-10-03 Hitachi Constr Mach Co Ltd 油圧ポンプ駆動系制御装置
JPS60222601A (ja) * 1984-04-20 1985-11-07 Komatsu Ltd 油圧制御装置
JPS614848A (ja) * 1984-06-18 1986-01-10 Hitachi Constr Mach Co Ltd 原動機と油圧ポンプを含む系の制御装置
JPS6111429A (ja) * 1984-06-26 1986-01-18 Hitachi Constr Mach Co Ltd 原動機と油圧ポンプを含む系の制御装置
GB2171757A (en) * 1985-02-28 1986-09-03 Komatsu Mfg Co Ltd I c engine driven variable displacement pumping systems
US4809504A (en) * 1986-01-11 1989-03-07 Hitachi Construction Machinery Co., Ltd. Control system for controlling input power to variable displacement hydraulic pumps of a hydraulic system
US4747335A (en) * 1986-12-22 1988-05-31 Caterpillar Inc. Load sensing circuit of load compensated direction control valve

Cited By (87)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US5083430A (en) * 1988-03-23 1992-01-28 Hitachi Construction Machinery Co., Ltd. Hydraulic driving apparatus
US5056312A (en) * 1988-07-08 1991-10-15 Hitachi Construction Machinery Co., Ltd. Hydraulic drive system for construction machines
US5025625A (en) * 1988-11-10 1991-06-25 Hitachi Construction Machinery Co., Ltd. Commonly housed directional and pressure compensation valves for load sensing control system
US5177964A (en) * 1989-01-27 1993-01-12 Hitachi Construction Machinery Co., Ltd. Hydraulic drive traveling system
US5101629A (en) * 1989-02-20 1992-04-07 Hitachi Construction Machinery Co., Ltd. Hydraulic circuit system for working machine
US5062350A (en) * 1989-03-22 1991-11-05 Hitachi Construction Machinery Co., Ltd. Hydraulic drive system for civil engineering and construction machine
US5079919A (en) * 1989-03-30 1992-01-14 Hitachi Construction Machinery Co., Ltd. Hydraulic drive system for crawler mounted vehicle
US5150574A (en) * 1989-05-02 1992-09-29 Hitachi Construction Machinery Co., Ltd. Hydraulic drive system for civil engineering and construction machine
US5077975A (en) * 1989-05-05 1992-01-07 Mannesmann Rexroth Gmbh Control for a load-dependently operating variable displacement pump
US5209063A (en) * 1989-05-24 1993-05-11 Kabushiki Kaisha Komatsu Seisakusho Hydraulic circuit utilizing a compensator pressure selecting value
US5170625A (en) * 1989-07-27 1992-12-15 Hitachi Construction Machinery Co., Ltd. Control system for hydraulic pump
US5152140A (en) * 1989-10-11 1992-10-06 Hitachi Construction Machinery Co., Ltd. Pressure compensating valve spool positioned by difference in pressure receiving areas for load and inlet pressures
US5101628A (en) * 1990-01-22 1992-04-07 Shin Caterpillar Mitsubishi Ltd. Energy regenerative circuit in a hydraulic apparatus
US5174114A (en) * 1990-02-28 1992-12-29 Hitachi Construction Machinery Co., Ltd. Hydraulic drive system for construction machine
US5129230A (en) * 1990-06-19 1992-07-14 Hitachi Construction Machinery Co., Ltd. Control system for load sensing hydraulic drive circuit
US5267440A (en) * 1990-09-11 1993-12-07 Hitachi Construction Machinery Co., Ltd. Hydraulic control system for construction machine
US5394696A (en) * 1990-12-15 1995-03-07 Barmag Ag Hydraulic system
US5297381A (en) * 1990-12-15 1994-03-29 Barmag Ag Hydraulic system
US5222417A (en) * 1991-01-22 1993-06-29 Fuji Jukogyo Kabushiki Kaisha Hydraulic control system of an automatic transmission for a motor vehicle
US5409038A (en) * 1991-01-23 1995-04-25 Kabushiki Kaisha Komatsu Seisakusho Hydraulic circuit including pressure compensating valve
US5307631A (en) * 1991-01-28 1994-05-03 Hitachi Construction Machinery Co., Ltd. Hydraulic control apparatus for hydraulic construction machine
AU642503B2 (en) * 1991-02-15 1993-10-21 Caterpillar Inc. Hydraulic circuit and control system therefor
WO1992014944A1 (en) * 1991-02-15 1992-09-03 Caterpillar Inc. Hydraulic circuit and control system therefor
US5138838A (en) * 1991-02-15 1992-08-18 Caterpillar Inc. Hydraulic circuit and control system therefor
US5351914A (en) * 1991-06-14 1994-10-04 Fuji Jukogyo Kabushiki Kaisha Hydraulic control system for aircraft
US5493861A (en) * 1991-08-19 1996-02-27 Danfoss A/S Hydraulic system with pump and load
US5267441A (en) * 1992-01-13 1993-12-07 Caterpillar Inc. Method and apparatus for limiting the power output of a hydraulic system
US5249421A (en) * 1992-01-13 1993-10-05 Caterpillar Inc. Hydraulic control apparatus with mode selection
US5207059A (en) * 1992-01-15 1993-05-04 Caterpillar Inc. Hydraulic control system having poppet and spool type valves
US5442912A (en) * 1992-12-04 1995-08-22 Hitachi Construction Machinery Co., Ltd. Hydraulic recovery device
US5447027A (en) * 1993-03-23 1995-09-05 Hitachi Construction Machinery Co., Ltd. Hydraulic drive system for hydraulic working machines
WO1994024407A1 (en) * 1993-04-19 1994-10-27 Bobbie Joe Bowden Automatic drilling system
US5501136A (en) * 1993-06-24 1996-03-26 Voac Hydraulics Boras Ab Control system for a hydraulic motor
US5659485A (en) * 1993-07-02 1997-08-19 Samsung Heavy Industry Co., Ltd. Discharge flow control system and method in hydraulic pump
US5540049A (en) * 1995-08-01 1996-07-30 Caterpillar Inc. Control system and method for a hydraulic actuator with velocity and force modulation control
US5896930A (en) * 1997-01-27 1999-04-27 Kabushiki Kaisha Kobe Seiko Sho Control system in hydraulic construction machine
US6209321B1 (en) * 1997-08-29 2001-04-03 Komatsu Ltd. Hydraulic controller for a working machine
US5930996A (en) * 1997-10-02 1999-08-03 Hitachi Construction Machinery Co., Ltd. Auto-acceleration system for prime mover of hydraulic construction machine and control system for prime mover and hydraulic pump
US6408622B1 (en) * 1998-12-28 2002-06-25 Hitachi Construction Machinery Co., Ltd. Hydraulic drive device
US6282891B1 (en) * 1999-10-19 2001-09-04 Caterpillar Inc. Method and system for controlling fluid flow in an electrohydraulic system having multiple hydraulic circuits
US6526747B2 (en) * 2000-01-25 2003-03-04 Hitachi Construction Machinery Co., Ltd. Hydraulic driving device
US6516932B2 (en) * 2000-09-29 2003-02-11 New Holland North America, Inc. Electro-hydraulic clutch hysteresis compensation
KR100415781B1 (ko) * 2000-10-06 2004-01-24 에스엠시 가부시키가이샤 자기센서가 부착된 전환밸브
US6701822B2 (en) 2001-10-12 2004-03-09 Caterpillar Inc Independent and regenerative mode fluid control system
US6715403B2 (en) 2001-10-12 2004-04-06 Caterpillar Inc Independent and regenerative mode fluid control system
US6684636B2 (en) * 2001-10-26 2004-02-03 Caterpillar Inc Electro-hydraulic pump control system
US6662705B2 (en) 2001-12-10 2003-12-16 Caterpillar Inc Electro-hydraulic valve control system and method
US7654337B2 (en) * 2002-11-05 2010-02-02 Sandvik Mining And Construction Oy Arrangement for controlling rock drilling
US20070007039A1 (en) * 2002-11-05 2007-01-11 Sandvik Tamrock Oy Arrangement for controlling rock drilling
US20060254671A1 (en) * 2002-12-09 2006-11-16 Endress + Hauser Flowtec Ag Method for filling a defined quantity of a medium into a container
US7458399B2 (en) * 2002-12-09 2008-12-02 Endress + Hauser Flowtec Ag Method for filling a defined quantity of a medium into a container
US6848254B2 (en) 2003-06-30 2005-02-01 Caterpillar Inc. Method and apparatus for controlling a hydraulic motor
US20040261407A1 (en) * 2003-06-30 2004-12-30 Hongliu Du Method and apparatus for controlling a hydraulic motor
US20050204735A1 (en) * 2004-03-17 2005-09-22 Kobelco Construction Machinery Co., Ltd. Hydraulic control system for working machine
US7392653B2 (en) * 2004-03-17 2008-07-01 Kobelco Construction Machinery Co., Ltd. Hydraulic control system for working machine
USRE44884E1 (en) 2004-09-24 2014-05-13 Stryker Corporation Ambulance cot with pinch safety feature
US8056950B2 (en) 2004-09-24 2011-11-15 Stryker Corporation In-ambulance cot shut-off device
US20090126361A1 (en) * 2005-11-25 2009-05-21 Hitachi Construction Machinery Co., Ltd Pump Torque Controller of Hydraulic Working Machine
US8056331B2 (en) * 2005-11-25 2011-11-15 Hitachi Construction Machinery Co., Ltd. Pump torque controller of hydraulic working machine
US20100218493A1 (en) * 2006-12-07 2010-09-02 Kazunori Nakamura Torque control apparatus for construction machine three-pump system
US8371117B2 (en) * 2006-12-07 2013-02-12 Hitachi Construction Machinery Co., Ltd. Torque control apparatus for construction machine three-pump system
US20110020146A1 (en) * 2008-03-31 2011-01-27 Teruo Akiyama Rotation drive controlling system for construction machine
US9022749B2 (en) * 2008-03-31 2015-05-05 Komatsu Ltd. Swing drive controlling system for construction machine
US9964965B2 (en) * 2009-12-10 2018-05-08 Hydraforce, Inc. Method of controlling proportional motion control valve
US20140299197A1 (en) * 2009-12-10 2014-10-09 Hydraforce, Inc. Method of controlling proportional motion control valve
US20120251332A1 (en) * 2009-12-24 2012-10-04 Doosan Infracore Co., Ltd. Power control apparatus and power control method of construction machine
KR20110073711A (ko) * 2009-12-24 2011-06-30 두산인프라코어 주식회사 건설기계의 동력제어장치
US8720629B2 (en) * 2009-12-24 2014-05-13 Doosan Infracore Co., Ltd. Power control apparatus and power control method of construction machine
KR101630457B1 (ko) 2009-12-24 2016-06-15 두산인프라코어 주식회사 건설기계의 동력제어장치
US20130121852A1 (en) * 2010-07-19 2013-05-16 Volvo Construction Equipment Ab System for controlling hydraulic pump in construction machine
US9303636B2 (en) * 2010-07-19 2016-04-05 Volvo Construction Equipment Ab System for controlling hydraulic pump in construction machine
US20130238178A1 (en) * 2012-03-07 2013-09-12 Clark Equipment Company Power management for a drive system
US9211808B2 (en) * 2012-03-07 2015-12-15 Clark Equipment Company Power management for a drive system
CN102607876A (zh) * 2012-04-13 2012-07-25 山东大学 适用于模型试验的多路高精度液压加卸载伺服控制系统
CN102607876B (zh) * 2012-04-13 2014-12-10 山东大学 适用于模型试验的多路高精度液压加卸载伺服控制系统
US9518655B2 (en) 2013-01-29 2016-12-13 Deere & Company Continuously adjustable control management for a hydraulic track system
US9145905B2 (en) 2013-03-15 2015-09-29 Oshkosh Corporation Independent load sensing for a vehicle hydraulic system
US20140366955A1 (en) * 2013-06-13 2014-12-18 Caterpillar Global Mining America Llc Remote regulator for roof bolter
US10343767B2 (en) * 2016-02-16 2019-07-09 The Boeing Company Hydraulic system and method for a flight control system of an aircraft
US10196131B2 (en) * 2016-02-16 2019-02-05 The Boeing Company Hydraulic system and method for an aircraft flight control system
CN107355437A (zh) * 2017-06-28 2017-11-17 安徽柳工起重机有限公司 负载敏感回转缓冲阀及起重机回转液压系统
US10989231B2 (en) * 2018-02-12 2021-04-27 Hawe Hydraulik Se Hydraulic valve assembly with forced circuit
US20210229924A1 (en) * 2018-07-06 2021-07-29 Sandvik Mining And Construction G.M.B.H. Hydraulic drive motor control system
US11827456B2 (en) * 2018-07-06 2023-11-28 Sandvik Mining And Construction G.M.B.H. Hydraulic drive motor control system
US11220804B2 (en) * 2019-07-26 2022-01-11 Robert Bosch Gmbh Hydraulic pressurizing medium supply assembly for a mobile work machine, and method
US11834811B2 (en) 2021-10-25 2023-12-05 Cnh Industrial America Llc System and method for controlling hydraulic pump operation within a work vehicle
US11608615B1 (en) * 2021-10-26 2023-03-21 Cnh Industrial America Llc System and method for controlling hydraulic valve operation within a work vehicle

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Publication number Publication date
EP0326150B1 (en) 1992-10-28
IN171213B (ko) 1992-08-15
DE68903281T2 (de) 1993-05-19
KR930002475B1 (ko) 1993-04-02
JPH01312202A (ja) 1989-12-18
DE68903281D1 (de) 1992-12-03
AU600400B2 (en) 1990-08-09
CN1035868A (zh) 1989-09-27
CN1010969B (zh) 1990-12-26
JPH07103881B2 (ja) 1995-11-08
EP0326150A1 (en) 1989-08-02
KR890012093A (ko) 1989-08-24
AU2886489A (en) 1989-07-27

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